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Steam Power Plant 
Engineering 



BY 

GEORGE F. GEBHARDT, M.E., A.M. 

PROrESSOR OF MECHANICAL ENGINEERING, ARMOUR 

INSTITUTE OF TECHNOLOGY 

CHICAGO, ILL. 



FIFTH EDITION, REWRITTEN AND RESET 

TOTAL ISSUE FIFTEEN THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 

London: CHAPMAN & HALL, Limited 

1917 



^" 



V^' 



T'J'400 



Copyright, 1908, 1910, 1913, 1917, 

BY 
G. F. GEBHARDT 



,. »^i"'f 



Stanbopc ipre«s 

F. H. GILSON COMPANY 
BOSTON. U.S.A. 



DEC 13 1917 



PREFACE TO FIFTH EDITION 



Although the first edition of this worlc was pubHshed less than a 
decade ago, the development of the Steam Power Plant has been so 
rapid that nearly all of the descriptive matter and a considerable portion 
of the data of this early edition became obsolete shortly after publica- 
tion. Revisions in 1909, 1911, and in 1913 failed to keep pace with the 
art, and the task of recording correct practice appeared to be a hopeless 
one. Fortunately radical changes during the past two years have been 
less marked and many of the elements entering into the modern Steam 
Power Plant have virtually reached the limit of efficiency, and some 
degree of stability may be expected from now on. It is quite unlikely 
that the Steam Power Plant of the immediate future will differ radically 
from the latest type already in operation, though increased boiler 
pressure, forced boiler capacity and the use of powdered low-grade fuel 
may effect minor changes. 

The same treatment of the subject has been followed in this edition 
as in earlier issues, but the book is to all intents and purposes a new one. 
Particular stress has been laid upon the subject of Fuels and Combustion 
and supplementary chapters on Elementary Thermodynamics, Proper- 
ties of Steam, and Properties of Dry and Saturated Air have been 
added at the request of practicing engineers. Numerous examples have 
been incorporated in the text, and the addition of typical exercises and 
problems may prove of value to the instructor. The scope of the work 
has been greatly enlarged, and with the exception of a few minor sections 
the entire text has been rewritten and reset. An extended study of 
certain portions of the work is facilitated by numerous references to 
current engineering literature. 

G. F. G. 

Chicago, Illinois, 
November, 1917. 



CONTENTS 



Page 

CHAPTER I. — Elementary Steam Power Plants 1-21 

1 . General 1 

2. Elementary Non-condensing Plant 2 

3. Non-condensing Plant; Exhaust Steam Heating 5 

4. Elementary Condensing Plant 7 

5. Condensing Plant with Full Complement of Heat-saving Devices. . . 13 

CHAPTER II. — Fuels and Combustion 22-109 

6. General 22 

7. Classification of Fuels 22 

8. Solid Fuels 22 

9. Composition of Coal 23 

10. Classification of Coals 30 

11. Anthracites 32 

12. Semi-anthracites 33 

13. Semi-bituminous 34 

14. Bituminous 35 

15. Sub-bituminous Coals 35 

16. Lignite or Brown Coals 37 

17. Peat or Turf 37 

18. Wood, Straw, Sawdust, Baggage, Tanbark 38 

19. Combustion 41 

20. Calorific Value of Coal 44 

21. Air Theoretically Required for Complete Combustion 47 

22. Products of Combustion 49 

23. Air Actually Supplied for Combustion 56 

24. Temperature of Combustion 58 

25. Heat Losses in Burning Coal 60 

26. Loss in the Dry Chimney Cases 61 

27. Loss Due to Incomplete Combustion 62 

28. Loss of Fuel through Grate 65 

29. Superheating the Moisture in the Air 66 

30. Loss Due to Moisture in the Fuel 66 

31. Loss Due to the Presence of Hydrogen in the Fuel 68 

32. Loss Due to Visible Smoke 68 

33. Radiation and Unaccounted for 68 

34. Heat Balance 69 

35. Standby Losses 71 

36. Inherent Losses 73 

37. Selection and Purchase of Coal 75 

38. Size of Coal — Bituminous 77 

39. Washed Coal 79 

V 



\d CONTENTS 

CHAPTER II — Continued Page 

40. Powdered Coal 80 

41. Types of Powdered Coal Feeders and Burners 81 

42. Boiler Furnaces for Burning Powdered Coal 84 

43. Cost of Preparing Powdered Coal 87 

44. Storing Powdered Coal 87 

45. Efficiency of Powdered Coal Furnaces 88 

46. Depreciation of Powdered Coal Furnaces 88 

47. Fuel Oil 89 

48. Chemical and Physical Properties of Fuel Oil 89 

49. Efficiency of Boilers with Fuel Oil 91 

50. Comparative Evaporative Economy of Oil and Coal 94 

51. Oil Burners 94 

52. Furnaces for Burning Fuel Oil 95 

53. Atomization of Oil , 101 

54. Oil-feeding Systems 101 

55. Oil Transportation and Storage 105 

56. The Purchase of Fuel Oil 108 

57. Gaseous Fuels 109 

CHAPTER III. — Boilers 115-181 

58. General 115 

59. Classification 115 

60. Vertical Tubular Boilers 116 

61. Fire-box Boilers 118 

62. Scotch-marine Boilers 119 

63. Robb-Mumford Boiler 121 

64. Horizontal Return Tubular Boilers 122 

65. Babcock & Wilcox Boilers 128 

66. Heine Boiler 130 

67. Parker Boiler 132 

68. Wickes Boiler . 135 

69. The Bigelow-Hornsby Boiler 135 

70. Stirling Boiler 136 

71. Winslow High-pressure Boiler 139 

72. Unit of Evaporation 143 

73. Heat Transmission 144 

74. Heating Surface 148 

75. The Horsepower of a Boiler 150 

76. Grate Surface and Rate of Combustion 151 

77. Boiler, Furnace and Grate Efficiency 154 

78. Boiler Capacity 163 

79. Effect of Capacity on Efficiency 165 

80. Economical Loads 168 

81. Influence of Initial Temperature on Economy 169 

82. Thickness of Fire 170 

83. Cost of Boilers and Settings 172 

84. Selection of Type 173 

85. Grates 174 

36. Shaking Grates 176 



CONTENTS vii 

CHAPTER 111 — Continued Page 

87. Blow-offs 177 

88. Damper Regulators 178 

89. Water Gauges 179 

90. Fusible or Safety Plugs 181 

91. Soot Blowers, Tube Cleaners, Etc 181 

CHAPTER IV. — Smoke Prevention, Furnaces, Stokers 187-222 

92. General 187 

93. Hand-fired Furnaces 190 

94. Dutch Ovens 192 

95. Twin-fire Furnace 194 

96. Chicago Settings for Hand-fired Return Tubular Boilers 194 

97. Burke's Smokeless Furnace 198 

98. Down-draft Furnaces 198 

99. Steam Jets 200 

rOO. Mechanical Stokers 201 

101. Chain Grates 203 

102. Overfeed Step Grates 207 

103. Underfeed Stokers 210 

104. Sprinkling 216 

105. Smoke Determination 217 

106. Cost of Stokers 222 

CHAPTER V. — Superheaters 223-243 

107. Advantages of Superheating 223 

108. Economy of Superheat 225 

109. Limit of Superheat 226 

110. Types of Superheaters 227 

111. Materials Used in Construction of Superheaters 236 

112. Extent of Superheating Surface 238 

113. Performance of Superheaters 243 

CHAPTER VI. — Coal and Ash-handling Apparatus 249-278 

114. General 249 

115. Coal Storage 249 

116. Coal-handhng Methods 250 

117. Hand Shoveling 251 

118. Continuous Conveyors and Elevators 252 

119. Elevating Tower, Hand-car Distribution 266 

120. Elevating Tower, Cable-car Distribution 268 

* 121. Hoist and Trolley; Telpherage 268 

122. Vacuum Conveyors 270 

123. Cost of Handhng Coal and Ashes 273 

124. Coal Hoppers 274 

125. Coal Valves 276 

CHAPTER VII. — Chimneys 279-326 

126. General 279 

127. Chimney Draft 280 

128. Chimney Area 291 



viii CONTENTS 

CHAPTER YU — Continued Page 

129. Empirical Chimney Equations 295 

130. Stacks for Oil Fuel 296 

131. Classification of Chimneys 286 

132. Guyed Chimneys 299 

133. Self-sustaining Steel Chimneys 300 

134. Wind Pressure 302 

135. Thickness of Plates for Self-sustaining Steel Stacks 303 

136. Riveting 305 

137. Stability of Steel Stacks 306 

138. Foundation Bolts for Steel Stacks 306 

139. Brick Chimneys 308 

140. Thickness of Walls 308 

141. Core and Lining 312 

142. Materials for Brick Chimneys 312 

143. Stability of Brick Chimneys 313 

144. Custodis Radial Brick Chimney 316 

145. Wiederholt Chimney 317 

146. Steel-concrete Chimneys 317 

147. Breeching 321 

148. Chimney Foundations 322 

149. Chimney Efficiencies 324 

150. Cost of Chimneys 325 

CHAPTER VIII. — Mechanical Draft 327-351 

151. General 327 

152. Steam Jets 328 

153. Fan Draft 330 

154. Types of Fans 337 

155. Performance of Fans 338 

156. Selection of Fan 345 

157. Chimney vs. Mechanical Draft 347 

CHAPTER IX. — Reciprocating Steam Engines 352-423 

158. Introductory 352 

159. The Ideal Engine 353 

160. Efficiency Standards 356 

161. Steam Consumption or Water Rate 357 

162. Heat Consumption : 358 

163. Thermal Efficiency 360 

164. Mechanical Efficiency 362 

165. Rankine Cycle Ratio 363 

166. Cylinder Efficiency 364 

167. Commercial Efficiencies 366 

168. Heat Losses in the Steam Engine 366 

169. Cylinder Condensation 366 

170. Leakage of Steam 369 

171. Clearance Volume 370 

172. Loss Due to Incomplete Expansion and Compression 371 

173. Loss Due to Wire Drawing 374 



CONTENTS ix 

CHAPTER IX — Continued Page 

174. Loss Due to Friction of the Mechanism 375 

175. Moisture 375 

176. Radiation and Minor Losses 376 

177. Heat Lost in the Exhaust 376 

178. Methods of Increasing Economy 380 

179. Increasing Boiler Pressure 380 

180. Increasing Rotative Speed 382 

181. Decreasing Back Pressure by Condensing 383 

182. Superheating 386 

183. Jackets 390 

184. Receiver Reheaters: Intermediate Reheating 391 

185. Compounding 392 

186. Unifiow or Unaflow Engine 393 

187. Use of Binary Vapors 398 

188. Types of Piston Engines 400 

189. High-speed Single-valve Simple Engines 400 

190. High-speed Multi- valve Simple Engines 403 

191. Medium and Low-speed Multi- valve Simple Engines 404 

192. Compound Engines 405 

193. Triple and Quadruple Expansion Engines 408 

194. The Locomobile 408 

195. Rotary Engines 410 

196. Throttling vs. Automatic Cut-Off 413 

197. Selection of Type 415 

198. Cost of Engines 416 

CHAPTER X. — Steam Turbines 424-500 

199. Classification 424 

200. General Elementary Theory 426 

201. The De Laval Turbine 430 

202. Elementary Theory — Single-wheel, Single-stage Turbine 432 

203. Terry Non-condensing Turbine 444 

204. Westinghouse Impulse Turbine 446 

205. Elementary Theory — Single-wheel, Multi-velocity-stage Turbine . . 446 
.206. De Laval Velocity-stage Turbine 447 

207. Kerr Turbine 448 

208. The De Laval Multi-stage Turbine 451 

209. Elementary Theory — Multi-pressure Single-velocity-stage Turbine. 452 

210. Terry Condensing Turbine 453 

211. Curtis Turbine 453 

212. Elementary Theory — Curtis Turbine 464 

213. Westinghouse Single-flow Reaction Turbine. . .'. 467 

214. Allis-Chalmers Steam Turbine 473 

215. Westinghouse Impulse-reaction Turbine 474 

216. Westinghouse Compound Steam Turbine 477 

217. Elementary Theory — Reaction Turbine 477 

218. Exhaust-steam Turbine — Low and Mixed Pressures 479 

219. Advantages of the Steam Turbine 486 

220. Efficiency and Economy of Steam Turbines 488 



X CONTENTS 

CHAPTER X — Continued Page 

221. Influence of Superheat 494 

222. Influence of High Initial Pressure 496 

223. Influence of High Vacua 496 

224. Tesla Bladeless Turbine 498 

225. ''Spiro" Turbines 498 

CHAPTER XI. — Condensers 501-566 

226. General 501 

227. Effect of Aqueous Vapor upon the Degree of Vacuum 505 

228. Gain in Power Due to Condensing 506 

229. Classification of Condensers 507 

230. Standard Low-level Jet Condensers 507 

231. Injection Orifice 511 

232. Volume of the Condenser Chamber 511 

233. Injection and Discharge Pipes 512 

234. High-vacuum Jet Condensers 512 

235. Siphon Condensers 514 

236. Size of Siphon Condensers 515 

237. Ejector Condenser 516 

238. Barometric Condenser 517 

239. Condensing Water, Jet Condensers 520 

240. Water-cooled Surface Condensers 523 

241. Coohng Water, Surface Condensers 527 

242. Heat Transmission through Condenser Tubes 529 

243. Dry-air Surface Condensers (Forced Circulation) 541 

244. Quantity of Air for Cooling (Dry-air Condenser) 542 

245. Saturated-air Surface Condensers (Natural Draft) 543 

246. Evaporative Surface Condenser 545 

247. Location and Arrangement of Condensers 547 

248. Cost of Condensers 553 

249. Choice of Condensers 556 

250. Water-cooling Systems 558 

251. Cooling Pond 558 

252. Spray Fountain 559 

253. Coohng Towers • 560 

254. Tests of Cooling Towers 563 

CHAPTER XII. — Feed Water Purifiers and Heaters 566-622 

255. General 566 

256. Scale 568 

257. Foaming and Priming 569 

258. Internal Corrosion 570 

259. General Feed Water Treatment 571 

260. Boiler Compounds 572 

261. Use of Kerosene and Petroleum Oils in Boiler Feed Water 573 

262. Use of Zinc in Boilers 573 

263. Methods of Introducing Compounds 573 

264. Mechanical Purification 574 

265. Thermal Purification 574 



CONTENTS xi 

CHAPTER Xll — Continued Page 

266. Water Softening 575 

267. Water-softening and Purifying Plants 577 

268. Economy of Preheating Feed Water 584 

269. Classification of Feed-water Heaters 585 

270. Open Heaters 586 

271. Combined Open Heater and Chemical Purifier 588 

272. Temperature in Open Heaters 588 

273. Pan Surface Required in Open Feed-water Heaters 589 

274. Size of Shell, Open Heaters 589 

275. Types of Closed Heaters 590 

276. Water-tube Closed Heaters 590 

277. Steam-tube Closed Heaters 593 

278. Film Heaters 594 

279. Heat Transmission in Closed Heaters 594 

280. Open vs. Closed Heaters 600 

281. Through Heaters 601 

282. Induced Heaters 602 

283. Live-steam Heaters and Purifiers 604 

284. Distillation of Make-up Water 606 

285. Fuel Economizers 607 

286. Temperature Rise in Economizers 611 

287. Value of Economizers 614 

288. Factors Determining Installation of Economizers 615 

289. Choice of Feed-water Heating System 616 

CHAPTER XIII. — Pumps 622-679 

290. Classification 622 

291. Boiler Feed Pumps, Direct-acting Duplex 624 

292. Feed Pumps with Steam-actuated Valves 627 

293. Air and Vacuum Chamber 628 

294. Water Pistons and Plungers 630 

295. Performance of Piston Pumps 632 

296. Size of Boiler-feed Pumps; Reciprocating Piston Type 638 

297. Steam Pump Governors 639 

298. Feed-water Regulators 640 

299. Power Pumps — Piston Type 644 

300. Injectors 645 

301. Positive Injectors 646 

302. Automatic Injectors 647 

303. Performance of Injectors 649 

304. Injectors vs. Steam Pump as a Boiler Feeder 650 

305. Vacuum Pumps 651 

306. Wet-air Pumps for Jet Condensers 652 

306-a. Wet-air Pumps for Surface Condensers 654 

307. Size of Wet-air Pumps 655 

308. Tail Pumps 657 

309. Dry-air or Dry Vacuum Pumps 658 

310. Size of Dry-air Pumps 661 

311. Centrifugal Pumps 664 



xii CONTENTS 

CHAPTER XlII — Continued Page 

312. Performance of Centrifugal Pumps 667 

313. Rotary Pumps 671 

314. Circulating Pumps 672 

315. Centrifugal Boiler-feed Pumps 674 

316. Condensate or Hot-well Pumps 675 

317. Air Lift 676 

CHAPTER XIV. — Separators, Traps, Drains 679-705 

318. Live-steam Separators; General 679 

319. Classification of Separators 680 

320. Tjrpes of Separators 681 

321. Location of Separators 684 

322. Exhaust-steam Separators and Oil Eliminators 685 

323. Exhaust Heads 687 

324. Drips 688 

325. Low-pressure Drips 688 

326. Size of Pipe for Low-pressure Drips 689 

327. High-pressure Drips 690 

328. Classification of Steam Traps 690 

329. Types of Traps 691 

330. Location of Traps 696 

331. Drips under Vacuum 698 

332. Drips under Alternate Pressure and Vacuum 699 

333. The Steam Loop 700 

334. The Holly Loop 701 

335. Returns Tank and Pump 703 

336. Office Building Drains 704 

CHAPTER XV. — Piping and Pipe Fittings 705-777 

337. General 705 

338. Drawings 705 

339. Materials for Pipes and Fittings 705 

340. Size and Strength of Commercial Pipe 707 

341. Screwed Fittings, Pipe Threads 709 

342. Flanged Fittings 711 

343. Loss of Heat from Bare and Covered Pipe 720 

344. Expansion 724 

345. Pipe Supports and Anchors 728 

346. General Arrangement of High-pressure Steam Piping 730 

347. Size of Steam Mains 734 

348. Flow of Steam in Pipes 740 

349. Friction through Valves and Fittings 745 

350. Equation of Pipes • 747 

351. Exhaust Piping, Condensing Plants 747 

352. Exhaust Piping, Non-condensing Plants, Webster Vacuum System. 749 

353. Exhaust Piping, Non-condensing Plants, Paul Heating System 751 

354. Automatic Temperature Control 752 

355. Feed-water Piping 754 

356. Flow of Water through Orifices, Nozzles and Pipes 757 



CONTENTS xiii 

CHAPTER XY — Continued Page 

357. Stop Valves . 761 

358. Automatic Non-return Valves 764 

359. Emergency Valves and Automatic Stops 765 

360. Check Valves 766 

361. Blow-off Cocks and Valves 767 

362. Safety Valves 768 

363. Back-pressure and Atmospheric Relief Valves '. 772 

364. Reducing Valves 774 

365. Foot Valves 775 

CHAPTER XVI. — Lubricants and Lubrication 777^804 

366. General 777 

367. Vegetable Oils 777 

368. Animal Fats 777 

369. Mineral Oils 778 

369-a. Solid Lubricants 779 

370. Greases 780 

371. Qualification of Good Lubricants 781 

372. Testing Lubricating Oils 781 

373. Chemical Tests of Lubricating Oils 781 

374. Physical Tests of Lubricating Oils 783 

375. Service Tests 787 

376. Atmospheric Surface Lubrication 789 

377. Intermittent Feed 789 

378. Restricted Feed 789 

379. Oil Bath 790 

380. Oil Cups 790 

381. Telescopic Oiler 791 

382. Ring Oiler 791 

383. Centrifugal Oiler 791 

384. Pendulum Oiler 791 

385. Splash Oiling 792 

386. Gravity Oil Feed 792 

387. Low-pressure Gravity Feed 793 

388. Compressed-air Feed 794 

389. Cylinder Lubrication 794 

390. Cylinder Cups 795 

391. Hydrostatic Lubricators 795 

392. Forced-feed Cylinder Lubrication 797 

393. Central Systems 798 

394. Oil Filters 801 

CHAPTER XVII. — Testing and Measuring Apparatus 804-843 

395. General 804 

396. Weighing the Fuel 804 

397. Measurement of Feed Water 806 

398. Actual Weighing of Feed Water 806 

399. Worthington Weight Determinator 806 

400. Kennicott Water Weigher 807 



xiv CONTENTS 

CHAPTER XYIl — Continued Page 

401. Wilcox Water Weigher 808 

402. Weir Measuring Devices 809 

403. Pressure Water Meters 809 

404. Venturi Meter 810 

405. Orifice Measurements 812 

406. Measurements of Steam 813 

407. Weighing Condensed Steam 813 

408. Steam Meters 813 

409. Pressure Gauges 825 

410. Measurement of Temperature 827 

411. Power Measurements 832 

412. Measurement of Speed 832 

413. Steam-engine Indicators 833 

414. Dynamometers 833 

415. Flue Gas Analysis 835 

416. Moisture in Steam 839 

417. Fuel Calorimeters 841 

418. Boiler Control Boards 843 



CHAPTER XVIII. — Finance and Economics — Cost of Power... 845-891 

419. General Records 845 

420. Permanent Statistics 847 

420-a. Operating Records 847 

421. Output and Load Factor 849 

422. Cost of Power. General 857 

423. Fixed Charges 858 

424. Interest 860 

425. Depreciation 861 

426. Maintenance 869 

427. Taxes and Insurance 869 

428. Operating Costs — General Division 869 

429. Labor, Attendance, Wages 871 

430. Cost of Fuel 872 

431. Oil, Waste and Supplies 877 

432. Repairs and Maintenance 877 

433. Cost of Power 878 

434. Elements of Power-plant Design 881 



CHAPTER XIX. — Typical Specifications 891-914 

435. Specifications for a Horizontal Tubular Boiler 891 

436. Specifications for Steam, Exhaust, Water and Condensing Piping 

for an Electric Power Station 898 

437. Government Specifications and Proposal for Supplying Coal 909 



CHAPTER XX. — Typical Central Stations 

438. Essex Station, Public Service, N.J 914-928 



CONTENTS XV 

Page 
CHAPTER XXI. — A Typical Modern Isolated Station 

439. Power Plant of the W. H. McElwain Company, Manchester, 

N. H 929-941 

CHAPTER XXII. — Supplementary — Properties of Saturated 

AND Superheated Steam 942-959 

440. General 942 

441. Notations 942 

442. Standard Units 942 

443. Quality 943 

444. Temperature-pressure Relation 943 

445. Specific Volume 943 

446. Heat of the Liquid 944 

447. Latent Heat of Vaporization 945 

448. Total Heat or Heat Content 946 

449. Specific Heat of Steam 947 

450. Entropy 951 

451. Molher Diagram 956 

CHAPTER XXIII. — Supplementary — Elementary Thermodynamics 

— Change of State 960 

452. General 960 

453. Isobaric or Equal Pressure Change 960 

454. Isovolumic or Equal Volume Change 962 

455. Isothermal or Equal Temperature Change 963 

456. Constant Heat Content 964 

457. Adiabatic Change of State 965 

458. Polytropic Change of State 967 

CHAPTER XXIV, — Supplementary — Elementary Thermodynamics 

OF the Steam Engine 971-990 

459. General 971 

460. Carnot Cycle 971 

461. Rankine Cycle; Complete Expansion 977 

462. Rankine Cycle with Incomplete Expansion 981 

463. Rankine Cycle with Rectangular PV-Diagram 982 

464. Conventional Diagram 983 

465. Logarithmic Diagram 985 

466. Temperature-entropy Diagram 987 

466-a. Steam Accounted for by Indicator Diagram d,t Points Near Cut- 

Off and Release 989 

CHAPTER XXV. — Supplementary — Properties of Air — Dry, 

Saturated, and Partially Saturated 991 

467. General 991 

* 468. Dry Air 991 

469. Saturated Air 994 

470. Partially Saturated Air 998 



xvi CONTENTS 

APPENDICES A-G 

Page 

APPENDIX A. — A.S.M.E. Boiler Testing Code 1007-1011 

APPENDIX B. — A.S.M.E. Engine Testing Code 1012-1016 

APPENDIX C. — A.S.M.E. Turbine Testing Code 1017-1019 

APPENDIX D. — A.S.M.E. Pumping Machinery Testing Code... 1020-1022 

APPENDIX E. — A.S.M.E. Power Plant Testing Code 1023-1030 

APPENDIX F. — Miscellaneous Conversion Factors 1031 

APPENDIX G. — Equivalent Values of Electrical and Me- 
chanical Units 1032 



STEAM POWER PLANT 
ENGINEERING 



CHAPTER I 

ELEMENTARY STEAM POWER PLANTS 

1. General. — By far the greater part of the mechanical and electrical 
energy generated for commercial purposes is furnished by the steam 
power plant. Despite the tremendous progress of the internal combus- 
tion engine and the rapid development of water power the steam plant 
is more than holding its own. 

The most efficient plant, thermally, in the conversion of energy 
from one form to another, is not necessarily the most economical com- 
mercially, since the various items involved in effecting this conversion 
may more than offset the gain over a less efficient plant. There is no 
question as to the low operating cost of power generated by hydro-elec- 
tric plants, but when the cost of transmission and the overhead charges 
are taken into consideration the economy is not so evident and may be 
completely neutralized. From a purely thermal standpoint the Diesel 
engine electric plant is superior to the best steam-electric plant for power 
purposes, but the fuel item is only one of the many involved in the total 
cost. It is the commercial efficiency which enables the steam power 
plant, with its extravagant waste of fuel, to compete successfully with 
the gas producer, internal-combustion engine and hydro-electric plant. 

A station which distributes power to a number of consumers more 
or less distant, is called a Central Station. When the distances are 
very great, electrical current of high tension is frequently employed, and 
is transformed and distributed at convenient points through Sub- 
stations: A plant designed to furnish power or heat to a building or a 
group of buildings under one management is called an Isolated Station. 
For example, the power plant of an office building is usually called an 
isolated station. 

When the exhaust steam from the engines is discharged at approxi- 
mately atmospheric pressure the plant is said to be operating non- 
condensing. W^hen the exhaust steam is condensed, reducing the back 

1 



2 STEAM POWER PLANT ENGINEERING 

pressure on the piston by the partial vacuum thus formed, the plant is 
said to operate condensing. 

When the exhaust steam may be used for manufacturing, heating, or 
other useful purposes, as is frequently the case in various manufacturing 
estabhshments, and in large office buildings, it is usually more eco- 
nomical to run non-condensing, while power plants for electric lighting 
and power, pumping stations, air-compressor plants, and others, in 
which the load is fairly constant and the exhaust steam is not required 
for heating, are generally operated condensing. 



stack 



Saiety 
Valve 




Steam Gi 



ooo o o o o 



Tubes - 



Throttle 



Engine 



Blow Off 



Boiler 



Fire 
Door 



o o o o 
ooooooo o 



/r~i 




O^::. 



Furnace 



Grate 



Ash Pit 



Injector 



Feed 

Water 

Tank 



Fig. 1. Elementary Non-condensing Plant. 



2. Elementary Non-condensing Plant. — Fig. 1 gives a diagrammatic 
outUne of the essential elements of the simplest form of steam power 
plant. The equipment is complete in every respect and embodies all 
the accessories necessary for successful operation. Where a small 
amount of power is desired at intermittent periods, as in hoisting 
systems, threshing outfits and traction machinery, the arrangement is 
substantially as illustrated. The output in these cases seldom exceeds 
50 horsepower and the time of operation is usually short, so the cheapest 
of appliances are installed, simplicity and low first cost being more im- 
portant than economy of fuel 



ELEMENTARY STEAM POWER PLANTS 3 

Such a plant has three essential elements: (1) The furnace, (2) the 
boiler, and (3) the engine. Fuel is fed into the furnace, where it is 
burned. A portion of the heat liberated from the fu(4 by combustion 
is absorbed by the water in the boiler, converting it into steam under 
pressure. The steam being admitted to the cylinder of the engine does 
work upon the piston and is then exhausted through a suitable pipe to 
the atmosphere. The process is a continuous one, fuel and water 
being fed into the furnace and the boiler in proportion to the power 
demanded. 

In such an elementary plant, certain accessories are necessary for 
successful operation. The grate for supporting the fuel during com- 
bustion consists of a cast-iron grid or of a number of cast-iron bars 
spaced in such a manner as to permit the passage of air through the 
fuel from below. The soUd waste products fall through or are ''sliced" 
through the grate bars into the ash pit, from which they may be re- 
moved through the ash door. The latter acts also as a means of regu- 
lating the supply of air below the grate. Fuel is fed into the furnace 
through the fire door, and when occasion demands, air may be supplied 
above the bed of fuel by means of this door. The combustion chamber 
is the space between the bed of fuel and the boiler heating surface, its 
office being to afford a space for the oxidation of the combustible gases 
from the solid fuel before they are cooled below ignition temperature by 
the comparatively cool surfaces of the boiler. The chimney or stack 
discharges the products of combustion into the atmosphere and serves 
to create the draft necessary to draw the air through the bed of fuel. 
Various forced-draft appliances are sometimes used to assist or to en- 
tirely replace the chimney. The heating surface is that portion of the 
boiler area which comes into contact with the hot furnace gases, absorbs 
the heat and transmits it to the water. In the small plant illustrated 
in Fig. 1, the major portion of the heating surface is composed of a num- 
ber of fire tubes below the Avater line, through which the heated gases 
pass. The superheating surface is that portion of the heating surface 
which is in contact with the heated gases of combustion on one side 
and steam on the other. The volume above the water level is called the 
steam space. Water is forced into the boilers either by a feed pu7np 
or an injector. In small plants of the type considered, steam pumps 
are seldom employed; the injector answers the purpose and is con- 
siderably cheaper. A safety valve connected to the steam space of the 
boiler automatically permits steam to escape to the atmosphere if an 
excessive pressure is reached. The water level is indicated by try 
cocks or by a gauge glass, the top of which is connected with the steam 
space and the bottom with the water space. Try cocks are small valves 



4 STEAM POWER PLANT ENGINEERING 

placed in the water column or boiler shell, one at normal water level, 
one above it, and one below. By opening the valves from time to time 
the water level is approximately ascertained. They are ordinarily 
used in case of accident to the gauge glass. Fusible plugs are frequently 
inserted in the boiler shell at the lowest permissible water level. They 
are composed of an alloy having a low fusing point which melts when 
in contact with steam, thus giving warning by the blast of the escaping 
steam if the water level gets dangerously low. The blow-off cock is a 
valve fitted to the lowest part of the boiler to drain it of water or to 
discharge the sediment which deposits in the bottom. The steam out- 
let of a boiler is usually called the steam 7iozzle. 

The essential accessories of the simple steam engine include: A 
throttle valve for controlling the supply of steam to the engine; the 
governor, which regulates the speed of the engine by governing the 
steam supply; the lubricator, attached to the steam pipe, which is 
usually of the '^sight-feed" class and provides for lubrication of piston 
and valve. Lubrication of the various bearings is effected by oil cups 
suitably located. Drips are placed wherever a water pocket is apt to 
form in order that the condensation may be drained. The apparatus 
to be driven by the engine may be direct connected to the crank shaft 
or belted to the flywheel or geared. 

In small plants of this type no attempt is made to utilize the exhaust 
steam except in instances where the stack is too short to create the 
necessary draft, in which case the exhaust may be discharged up the 
stack. If the draft is produced by convection of the heated gases in 
the chimney, the fuel is said to be burned under natural draft; if the 
natural draft is assisted by the exhaust steam, the fuel is said to be 
burned under forced draft. The power realized from a given weight of 
fuel is very low and seldom exceeds 2J per cent of the heat value of the 
fuel. The distribution of the various losses in a plant of, say, 40 horse 
power is approximately as follows: 

B.t.u. 

Heat value of 1 pound of coal 14,500 

Boiler and furnace losses, 50 per cent 7,250 

Heat equivalent of one horsepower-hour 2,546 

Heat used to develop one horsepower-hour (50 pounds steam 
per horsepower-hour, pressure 80 pounds gauge, feed water 

67 deg. fahr.) 57,500 

2 546 ^^'^ ''^''*- 

Percentage of heat of the steam realized as work, ' 4.4 

o ( ,oUU 

Percentage of heat value of the coal realized as work, .,- _ ' r-— 2.2 

o7,oU0 -^ U.oU 

In Europe small non-condensing plants are developed to a high 
degree of efficiency. Through the use of highly superheated steam, 



ELEMENTARY STEAM POWER PLANTS 5 

specially designed engines and boilers, plants of this type as small as 
40 horsepower are operated with over-all efficiencies of from 10 to 12 
per cent. 

The power plant of the modern locomotive is very much like that 
illustrated in Fig. 1, the main difference lying in the type of boiler and 
engine. The entire exhaust from the engine is discharged up the 
stack through a suitable nozzle, since the extreme rate of combustion 
requires an intense draft. The engine is a highly efficient one com- 
pared with that in the illustration, and the performance of the boiler is 
more economical. In average locomotive practice about 5 per cent of 
the heat value of the fuel is converted into mechanical energy at the 
draw bar.* In general, a non-condensing steam plant in which the 
heat of the exhaust is wasted is very uneconomical of fuel, even under 
the most favorable conditions, and seldom transforms as much as 7 per 
cent of the heat value of the fuel into mechanical energy. 

3. Non-condensing Plant. Exhaust Steam Heating. — Fig. 2 gives a 
diagrammatic arrangement of a simple non-condensing plant differ- 
ing from Fig. 1 in that the exhaust steam is used for heating purposes. 
This shows the essential elements and accessories, but omits a number 
of small valves, by-passes, drains, and the like for the sake of sim- 
plicity. The plant is assumed to be of sufficient size to warrant the 
installation of efficient appUances. Steam is led from the boiler to 
the engine by the steam main. The moisture is removed from the 
steam before it enters the cylinder by a steam separator. The moisture 
drained from the separator is either discharged to waste or returned to 
the boiler.. The exhaust steam from the engine is discharged into the 
exhaust main where it mingles with the steam exhausted from the steam 
pumps. Since the exhaust from engines and pumps contains a large 
portion of the cylinder oil introduced into the live steam for lubricating 
purposes, it passes through an oil separator before entering the heat- 
ing system. After leaving the oil separator the exhaust steam is di- 
verted into two paths, part of it entering the feed-water heater where 
it condenses and gives up heat to the feed water, and the remainder 
flowing to the heating system. During warm weather the engine 
generally exhausts more steam than is necessary for heating purposes, 
in which case the surplus steam is automatically discharged to the 
exhaust head through the hack-pressure valve. The back-pressure valve 
is, virtually, a large weighted check valve which remains closed when 
the pressure in the heating system is below a certain prescribed amount, 
but which opens automatically when the pressure is greater than this 
amount. During cold weather it often happens that the engine ex- 
* Best modern practice gives about 8 per cent as a maximum. 



STEAM POWER PLANT ENGINEERING 




ELEMENTARY STEAM POWER PLANTS 7 

haust is insufficient to supply the heating system, the radiators con- 
densing the steam more rapidly than it can be suppHed. In this case 
live steam from the boiler is automatically fed into the main heating 
supply pipe through the reducing valve. 

The condensed steam and the entrained air which is always present 
are automatically discharged from the radiators by a thermostatic valve 
into the returns header. The thermostatic valve is so constructed that 
when in contact with the comparatively cool water of condensation it 
remains open and when in contact with steam it closes. The vacuum 
pump or vapor pump exhausts the condensed steam and air from the 
returns header and discharges them to the returns tank. The small 
pipe *S admits cold water to the vacuum pump and serves to condense 
the heated vapor, and at the same time supply the necessary make-up 
water to the system. The returns tank is open to the atmosphere so 
that the air discharged from the vacuum pump may escape. From 
the returns tank the condensed steam gravitates to the feed-water 
heater where its temperature is raised to practically that of the exhaust 
steam. The feed water gravitates to the feed pump and is forced into 
the boiler. There are several systems of exhaust steam heating in cur- 
rent practice which differ considerably in details, but, in a broad sense, 
are similar to the one just described. The more important of these 
will be described later on. 

During the summer months when the heating system is shut down, 
the plant operates as a simple non-condensing station and practically 
all of the exhaust steam, amounting to perhaps 80 per cent of the heat 
value of the fuel, is wasted. The total coal consumption, therefore, is 
charged against the power developed. During the winter months, 
however, all, or nearly all, of the exhaust steam may be used for heating 
purposes and the power becomes a relatively small percentage of the 
total fuel energy utiUzed. The percentage of heat value of the fuel 
chargeable to power depends upon the size of the plant, the number 
and character of engines and boilers, and the conditions of operation. 
It ranges anywhere from 50 to 100 per cent for the summer months 
and may run as low as 6 per cent for the winter months. This is on the 
assumption, of course, that the engine is debited only with the differ- 
ence between the coal necessary to produce the heat entering the cyl- 
inder and that utilized in the heating system. 

4. Elementary Condensing Plant. — Under the most favorable con- 
ditions a non-condensing plant cannot be expected to reaUze more 
than 10 per cent of the heat value of the fuel as power. In large non- 
condensing power stations the demand for exhaust steam is usually 
limited to the heating of the feed water, and as only 12 or 15 per cent 



8 STEAM POWER PLANT ENGINEERING 

can be utilized in this manner, the greater portion of the heat in the 
exhaust is lost. Non-condensing engines using saturated steam require 
from 20 to 60 pounds of steam per hour for each horsepower developed. 
On the other hand in condensing engines the steam consumption may 
be reduced to as low as 10 pounds per horsepower-hour. The saving of 
fuel is at once apparent. 

Fig. 3 gives a diagrammatic arrangement of a simple condensing plant 
in which the back pressure on the engine is reduced by condensing 
the exhaust steam. A different type of boiler from that in Fig. 1 or 
Fig. 2 has been selected, for the purpose of bringing out a few of the 
characteristic elements. The products of combustion instead of pass- 
ing directly through fire tubes to the stack as in Fig. 1 are deflected 
back and forth across a number of water tubes, by the bridge wall and a 
series of baffles. After imparting the greater part of their heat to the 
heating surface the products of combustion escape to the chimney 
through the breeching or flue. The rate of flow is regulated by a damper 
placed in the breeching as indicated. 

The steam generated in the boiler is led to the engine through the 
main header. The steam is exhausted into a condenser in which its 
latent heat is absorbed by injection or cooling water. The process 
condenses the steam and creates a partial vacuum. The condensed 
steam, injection water, and the air which is invariably present are with- 
drawn by an air pump and discharged to the hot well. In case the 
vacuum should fail, as by stoppage of the air pump, the exhaust steam 
is automatically discharged to the exhaust head by the atmospheric 
relief valve, and the engine will operate non-condensing. The atmos- 
pheric relief valve is a large check valve which is held closed by atmos- 
pheric pressure as long as there is a vacuum in the condenser. When 
the vacuum fails the pressure of the exhaust becomes greater than that 
of the atmosphere and the valve opens. 

The feed water may be taken from the hot well or from any other 
source of supply and forced into the heater. In this particular case it is 
taken from a cold supply and upon entering the heater is heated by the 
exhaust steam from the air and feed pumps. From the heater it gravi- 
tates to the feed pump and is forced into the boiler. Various other 
combinations of heaters, pumps, and condensers are necessary in many 
cases, depending upon the conditions of operation. Feed pumps, air 
pumps, and in fact all small engines used in connection with a steam 
power plant are usually called auxiliaries. 

A well-designed station similar to the one illustrated in Fig. 3 is 
capable of converting about 12 per cent of the heat value of the fuel 



ELEMENTARY STEAM POWER PLANTS 



9 




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10 STEAM POWER PLANT ENGINEERING 

into mechanical energy. The various heat losses under average con- 
ditions are approximately as follows: 

BOILER LOSSES. 

Per cent. 

Loss due to fuel falling through the grate 4 

Loss due to incomplete combustion 2 

Loss to heat carried away in chimney gases 20 

Radiation and other losses 9 

Total 35 

Heat used by engines and auxiliaries (16 pounds of steam per B.t.u. 

i.h.p-hour, pressure 150 pounds, feed water 210 deg. fahr 16,250 

Engine and generator friction, 5 per cent 812 

Leakage, radiation, etc., 2 per cent 325 

Total 17,387 

Heat equivalent of one electrical horsepower 2,546 

Percentage of the heat value of the steam converted into electrical Per cent. 

energy 14.7 

Percentage of heat value of fuel converted into electrical energy 

2546 X 0.65 

17,387 ^-^ 

One of the best recorded American performances of a reciprocating 
engine steam-electric power plant is that of the Pacific Light and Power 
Company at Redondo, Cal. When operating under regular commer- 
cial conditions approximately 14 per cent of the available heat of the 
fuel (crude oil) is realized as power at the switchboard. This includes 
all standby losses. For a detailed description of the plant and the 
results of the acceptance tests, see Jour, of Elec. Gas and Power, Aug. 
22, 1908. 

Fig. 4 gives a diagrammatic arrangement of one section of a modern 
large turbo-alternator central station. Each section is to all intents 
and purposes an independent plant. It will be noted that the essen- 
tial elements are practically the same as in the reciprocating station 
engine plant. Fig. 3, differing only in size and design. 

The power house, coal storage pile, storage and switch tracks, over- 
head bunkers, and coal and ash conveyors have been omitted for the 
sake of simpHcity, though the fuel supply and distributing system is 
an important factor in the design and operation of the plant. In the 
very latest designs the entire coal and ash handling equipment is elec- 
trically operated from a centralized board control. Assuming the coal 
bunkers over the boilers to be supplied with fuel, the operation is as 
follows : Coal descends by gravity to the stokers which, in this particular 
case, are of the under-feed, sloping fire-bed type. Ash and clinkers 
are removed by cUnker grinders located in a pit and are discharged into 
the ash hopper. Motor-driven blowers supply the air required for com- 



ELEMENTARY STEAM POWER PLANTS 



11 




12 STEAM POWER PLANT ENGINEERING 

bustion. Tliis air, in some installations, is preheated by the recovery of 
radiation and electrical losses in the turbine room and by radiation from 
the steam pipes. 

The boilers are much larger individually and fewer in number than 
in the old style reciprocating-engine plant and generate steam at 250 
to 300 pounds pressure, superheated to approximately 650 deg. fahr. 
When operating the turbines at full load the boilers are driven at 175 to 
200 per cent or more of their commercial rating. Reserve or spare 
boilers are conspicuous by their absence. When a boiler is cut out for 
repairs the rest of the battery is operated at from 225 to 275 per cent 
rating or more in order to evaporate the required amount of water. 
Each battery is designed to furnish steam directly to one particular 
turbine but by means of a cross-over main the steam from any battery 
of boilers may flow to any turbine. 

The prijne movers are horizontal steam turbines direct connected to 
alternators. The bearings are water cooled and lubrication is auto- 
matically effected by means of a pump connected to the governor shaft. 
Each turbine is normally excited by the main exciter mounted on an 
extension of the generator shaft. The generator field may also be ex- 
cited from an independently driven exciter or from the station storage 
battery. Air, washed and conditioned if necessary, is drawn into the 
generator by the revolving member and absorbs the electrical heat losses. 
The efficiency of the generator is very high (96 per cent) and yet be- 
cause of the great amount of energy transformed in the generator this 
4 per cent loss represents a large amount of heat and forced ventilation 
is necessary to prevent overheating. This preheated air may be dis- 
charged to waste or carried through conduits to the forced draft fan. 

The condenser is ordinarily of the surface condensing type and is 
attached directly below the low pressure end of the turbine. A much 
higher vacuum is maintained in the condensers than in reciprocating 
engine practice since the turbine gives its best efficiency at low back 
pressures. Condensing water is circulated through the tubes of the 
condenser by motor-driven or steam-driven centrifugal pumps and the 
condensed steam or condensate collected in the hot wells is withdrawn 
by a turbine-driven or motor-driven hot-well pump. Air and non-con- 
densable vapors are removed by a reciprocating dry air pump, steam or 
electrically driven. Rotary air pumps and so called turbo-air pumps 
are also used for this purpose. The hot-well pump discharges the con- 
4ensate into a feed-water heater which receives the steam exhausted 
from the steam driven auxiliaries. The steam turbine-driven centrif- 
ugal boiler feed pump takes its supply from the feed-water heater and 
delivers it to the boiler. 



ELEMENTARY STEAM POWER PLANTS 13 

A modern station similar to the one illustrated in Fig. 4, equipped with 
30,000 kilowatt units, is capable of converting over 18 per cent of the 
heat value of the fuel into electrical energy when operating at its most 
economical load. Under commercial conditions of operation with its 
attendant standby losses the average overall efficiency ranges from 12 
to 16 per cent. 

5. Condensing Plant with Full Complement of Heat-saving Appli- 
ances. — When fuel is costly it frequently becomes necessary for the 
sake of economy to reduce the heat wastes as much as possible. The 
chimney gases, w^hich in average practice are discharged at a tempera- 
ture between 450 and 550 deg. fahr., represent a loss of 20 to 30 per 
cent of the total value of the fuel. If part of the heat could be re- 
claimed without impairing the draft the gain would be directly propor- 
tional to the reduction in temperature of the gases. Again, in some 
types of condensers all of the steam exhausted by the engine is con- 
densed by the circulating water and discharged to waste. If provi- 
sion could be made for utilizing part of the exhaust steam for feed-water 
heating, the efficiency of the plant could be correspondingly increased. 
In many cases the cost of installing such heat-saving devices would 
more than offset the gain effected, but occasions arise where they give 
marked economy. 

Fig. 5 gives a diagrammatic arrangement of a condensing plant in 
which a number of heat-reclaiming devices are installed. The plant is 
assumed to consist of a number of engines, boilers, and auxiliaries. 
Coal is automatically transferred from the cars to coal hoppers placed 
above the boiler, by a system of buckets and conveyors. These hoppers 
store the coal in sufficient quantities to keep the boiler in continuous 
operation for some time. From the hoppers the coal is fed inter- 
mittently to the stoker by means of a down spout. The stoker feeds the 
furnace in proportion to the power demanded and automatically re- 
jects the ash and refuse to the ash pit. The ashes are removed from the 
ash pit, when occasion demands, and are transferred to the ash hop- 
per by the same system of buckets and conveyor which handles the coal. 
The ash hopper is usually placed alongside the coal hoppers and is not 
unlike them in general appearance and construction. 

The products of combustion are discharged to the stack through the 
flue or breeching. Within the flue is placed a feed-water heater called 
an economizer, the function of which is to absorb part of the heat from 
the gases on their way to the chimney. The heat reclaimed by the 
economizer- varies widely with the conditions of operation and ranges 
between 5 and 20 per cent. Since the economizer acts as a resistance 
to the passage of the products of combustion it is sometimes necessary 



14 



STEAM POWER PLANT ENGINEERING 




M 

o 



P5 






i 

I 



•I 



I 

I 

i 



to increase the draft either by increasing the height of the chimney 
or, as is the usual practice, by using a forced-draft system. 

Part of the heat of the exhaust steam is reclaimed by a vacuum heater 
which is placed in the exhaust line between engine and condenser. 
For example, if the feed water has a normal temperature of 60 deg. f ahr. 
and the vacuum in the condenser is 26 inches, the vacuum heater will 



ELEMENTARY STEAM POWER PLANTS 15 

raise the temperature of the feed to, say, 120 deg. fahr., thereby effect- 
ing a gain in heat of approximately 6 per cent. If the feed supply is 
taken from the hot well the vacuum heater is without purpose, as the 
temperature of the hot well will not be far from 120 deg. fahr. 

Referring to the diagram, the path of the steam is as follows: From 
the boiler it flows through the boiler lead to the ynain header or equaliz- 
ing pipe. From the main header it flows through the engine lead to the 
high-pressure cylinder. The exhaust steam discharges from the low- 
pressure cylinder through the vacuum heater and into the condenser. 
Part of the exhaust steam is condensed in the vacuum heater and gives 
up its latent heat to the feed water. The remainder is condensed by 
the injection water which is forced into the condenser chamber by the 
circulating pump. The condensed steam and circulating water gravitate 
through the tail pipe to the hot well. The air which enters the condenser 
either as leakage or entrainment is withdrawn by the air pump. The 
steam exhausted by the feed pump, air pump, stoker engine, and other 
steam-driven auxiharies is usually discharged into the atmospheric 
heater, which still further heats the feed water. 

Referring to the feed water, the circuit is as follows: The pump draws 
in cold water at a temperature of, say, 60 deg. fahr., and forces it in 
turn through the vacuum heater, the atmospheric heater, and the 
economizer into the boiler. The vacuum heater raises the temperature 
to 120 deg. fahr., the atmospheric heater increases it to 212 deg. fahr., 
and the economizer still further to about 300 deg. fahr. The heat 
reclaimed by this series of heaters is evidently the equivalent of that 
necessary to raise the feed water from 60 deg. fahr. to 300 deg. fahr., 
or approximately 24 per cent of the total steam supplied. In some 
plants the economizer only is installed; in others the economizer and 
atmospheric heater are deemed desirable; still others utilize all three. 
The distribution of the heat losses in a plant of this type using saturated 
steam and operating under favorable conditions is approximately as 
follows : 

Per Cent. B.t.u. 
Delivered to engine, 15 pounds steam per i.hp-hour; 

pressure 150 pounds, feed 60 deg. fahr 100 17,482 

Pelivered to feed pump 1.5 262 

Delivered to circulating pump 1.5 262 

Delivered to air pump 2 349 

Delivered to small auxiliaries 1.5 262 

Loss in leakage and drips 0.5 87 

Engine and generator friction 5 874 

Radiation and minor losses 1 175 

Total 19,753 



16 STEAM POWER PLANT ENGINEERING 

Per Cent. B.t.u. 

Returned by vacuum heater 5.5 1,086 

Returned by atmospheric heater 7.9 1,560 

Returned by economizer 9.7 1,916 

Total 23. 1 4,562 

Net heat dehvered to engine in the form of steam to pro- 
duce one electrical horsepower, 19,753 — 4,562 15,191 

Percentage converted to electrical power ' 16. 7 

Boiler efficiency 70 

Percentage of heat value of fuel necessary to produce one - 

electrical horsepower at switchboard — .... 11.7 

The preceding figures refer to reciprocating engine plants only and 
give the results of very good practice. So much depends upon the 
size and character of the prime movers, the nature of the fuel, and 
the conditions of operation that no definite figure can be given for the 
percentage of heat converted to power in a given type of station. Six 
per cent represents good average practice in a non-condensing plant 
and 10 per cent in a condensing plant using saturated steam. Pumping 
stations operating continuously under full load have reahzed as much as 
15 per cent of the total heat value of the fuel, but such performances 
are practically unobtainable in connection with reciprocating engine 
steam-driven electrical power plants with the usual peak loads and low 
yearly load factor. 

Figure 6 gives a diagrammatic arrangement of the essential elements 
of a modern steam turbine plant including the various heat-reclaiming 
devices described in the preceding paragraph. Turbo-generator No. 3 
of the Northwest Station, Commonwealth Edison Company, Chicago, 
Illinois, is a well-known example of this arrangement of prime mover 
and auxiliaries. In the Northwest Station the prime mover is a hori- 
zontal turbine-generator of 30,000 kilowatt rated capacity at 100 per 
cent power factor, with high and low pressure cyhnders mounted in 
tandem on the same shaft. The exciter is direct connected to the main 
generator and is rated at 110 kilowatts. Approximately 60,000 cubic 
feet per minute of cooling air are required to ventilate the generator 
and carry away the electrical heat losses. All bearings are water cooled 
and the oil supply is automatically maintained by a pump connected to 
the shaft of the governor. The boilers, five in number, are rated at 
1220 horsepower each, and are equipped with traveUng chain grates. 
When operating at full capacity they are capable of delivering about 
400,000 pounds of steam per hour. Steam is generated at a pressure of 
250 pounds gauge and superheated to a final temperature of approxi- 



ELEMENTARY STEAM POWER PLANTS 



17 




18 STEAM POWER PLANT ENGINEERING 

mately 625 deg. fahr. The main header is 20 inches in diameter and 
delivers the steam to the turbine at a velocity of about 8000 feet per 
minute at rated load. The condenser is of the surface type, contains 
50,000 square feet of cooling surface and is attached directly below the 
low pressure turbine. Circulating water is obtained from the river 
through an intake tunnel fitted with revolving screens. A double 
suction centrifugal pump of 52,000 gallons per minute capacity delivers 
-the water to the condenser against a static head of 15 feet. This pump 
is driven by a 650 horsepower non-condensing turbine at 1500 r.p.m. 
In passing through the condenser the temperature of the circulating 
water is raised approximately 15 degrees. The air and condensate are 
withdrawn from the condenser through a single pipe which connects 
with the separating chamber of a combination condensate and air pump 
of the "turbo-air" type. This pump is mounted on the same shaft 
with the turbine-driven circulating pump. The condensate at a tem- 
perature of 85 deg. fahr. is removed from the separating chamber by the 
'^condensate" end of the combination pump and delivered to the pre- 
heater or vacuum heater. The air is removed by. the "hurling water" 
end of the combination pump and delivered to the hurhng water reser- 
voir. The hurhng water is used over and over again aiid serves only 
for the ejection of air. 

' The condensate in passing through the preheater, which is located 
in the condenser near the opening of the exhaust of the low pressure 
turbine, has its temperature raised to 100 deg. fahr. After passing 
through the preheater the condensate is discharged into the atmospheric 
heater where its temperature is increased to 160 deg. fahr. by the 
exhaust steam from the steam-driven auxiharies. The overflow from 
the heater is discharged into the hot water reservoir from which a certain 
amount is drawn into the condenser through the agency of an auto- 
matically controlled float located in the heater. This system of drawing 
make-up and overflow water into the condenser becomes inoperative 
when the distance through which the water is to be lifted is sufficiently 
great to cause "vapor binding." The higher the temperature of the 
water in the reservoir the lower will be the permissible lift. The ob- 
ject of this arrangement is to maintain a continuous supply to the boilers 
irrespective of the fluctuation in the amount of condensate and to con- 
serve the overflow. The boiler feed pumps are turbine-driven three- 
stage centrifugal pumps, and at a speed of 2500 r.p.m. are capable of 
delivering 400,000 pounds of condensate into the economizers at a 
pressure of 315 pounds gauge. Each boiler has its own economizer 
and independently driven induced draft fan. Each economizer con- 
tains 7300 square feet of heating surface and is inserted between the 



ELEMENTARY STEAM POWER PLANTS 19 

lx)iler and the breaching. The feed water is raised from a temperature 
of 160 deg. fahr. to 270 dcg. fahr. in passing through the economizers. 
The five 100 horsepower motor-driven induced-draft fans maintain a 
draft in the boiler uptakes of 2.4 inches and are capable of handUng 
90,000 cubic feet of gases per minute. 

The thermal efficiency of these very large steam turbines is higher 
than that of the gas engine and is excelled only by engines of the Diesel 
type. Allowing an average steam consumption of 12 pounds per kilo- 
watt-hour for the turbine and all its auxiliaries and a boiler and econo- 
mizer efficiency of 80 per cent, the over-all efficiency from switchboard 
to coal pile is approximately 20 per cent. The over-all efficiency meas- 
ured over the period of one year is somewhat less than this because of 
great variation in load and the accompanying standby losses. 

In Europe a combined plant efficiency of 15 per cent is not uncom- 
mon. Even small semi-portable plants of 40 horsepower are operated 
with over-all efficiencies as high as 14 per cent. In these small plants 
the engine, boiler, and auxiliaries are combined, permitting a high degree 
of superheat with minimum heat losses. A 40-horsepower plant tested 
by Professor Josse of the Royal Technical School, Germany, gave the 
following results: coal consumed per brake hp-hour, 1.23 pounds, 
corresponding to an over-all efficiency of 14.2 per cent. Steam con- 
sumption, 9.5 pounds per i.hp-hour. Boiler and superheated efficiency, 
77.7 per cent. (See Zeit. des Ver. Deut. Ingr., March 18 and 25, 1911, 
and Power, Sept. 27, 1910, p. 1714.) 

The remarkable economy which i^ being effected in Europe with 
this type of plant is still further marked by the performance of a 100 
horsepower Wolf tandem compound locomobile which is credited with 
a performance of one brake horsepower per 0.86 pound of coal, corre- 
sponding to an over-all efficiency of 20 per cent. (Zeit. des Ver. Deut. 
Ingr., June, 1911.) 

The percentage of the heat value of the fuel realized as energy at 
the point of consumption is considerably less than the over-all efficiency 
from ''coal-pile to switchboard" because of the transmission, dis- 
tribution and service losses. These losses vary within wide limits, 
depending upon the size and type of plant, character of equipment, 
length of transmission lines and various other influencing factors. 
Figure 7 illustrates the approximate losses for a large plant such as the 
Northwest Station of the Commonwealth Edison Company of Chicago. 

For a description of the Ford Gas-Steam Plant see Power, Nov. 21, 
1916, p. 706; Jan. 16, 1917, p. 70. 



20 



STEAM POWER PLANT ENGINEERING 




ELEMENTARY STEAM POWER PLANTS 21 



EXERCISES. 

1. Make a diagrammatic outline of a simple non-condensing plant correctly 
locating all the essential elements entering into its composition. Indicate by means 
of arrow points the direction of flow of the feed water and steam. 

2. Same instruction as in Problem 1, except that a non-condensing plant with 
exhaust steam heating system is to be considered. 

3. Enumerate the character and extent of the heat losses from " coal-pile to switch- 
board" in a simple non-condensing piston engine plant. 

4. Beginning with the cold water supply trace the path of the feed water and steam 
through the various essential elements in a condensing plant equipped with a full 
complement of "heat-saving" appliances. 

5. Make a skeleton outline of a modern turbo-alternator plant correctly locating 
and designating by name all the essential elements entering into its composition. 



CHAPTER II 

FUELS AND COMBUSTION 

6. General. — The cost of fuel is by far the greatest single item of 
expense in the production of power and ranges from 40 per cent to 70 
per cent of the total operating costs. Furthermore, all fuels are slowly 
but surely increasing in price and larger investments for fuel saving 
equipment are justified. In localities where a specific fuel is plentiful 
the problem resolves itself merely into a study of the best methods of 
burning this fuel, but in situations where various kinds of fuel are avail- 
able the selection of the one best suited for a given or proposed equipment 
includes a careful consideration of such items as composition of the fuel, 
size, cost per ton, heating value, refuse incident to combustion, initial 
waste products such as ash and moisture, storage requirements, and 
transportation facilities. 

The fuels used for steam making are coal, coke, wood, peat, mineral 
oil, natural and artificial gases, refuse products such as straw, manure, 
sawdust, tan bark, bagasse, and garbage. 

In most cases that fuel is selected which develops the required power 
at the lowest cost, taking into consideration all of the circumstances 
that may affect its use. Occasionally the disposition of waste products 
is a factor in the clu)ice, but such instances are uncommon. The 
boilers and furnaces are designed to suit the fuel selected. 

7. Classification of Fuels. — Fuels may be divided into three classes 
as follows: 

1. Solid fuels. 

a. Natural: straw, wood, peat, coal. 

h. Prepared: charcoal, coke, peat, and briquetted fuels. 

2. Liquid fuels. 

a. Natural: crude oils. 
h. Prepared: distilled oils. 

3. Gaseous fuels. 

a. Natural : natural gas. 

6. Prepared: coal gas, water gas, producer gas, oil gas. 

8. Solid Fuels. — Solid fuels are of vegetable origin and exist in a 
variety of forms between that of a comparatively recent cellulose growth 

22 



FUELS AND COMBUSTION 



23 



and that of nearly pure carbon as anthracite coal. They owe their 
forms to the conditions under which they were created or to the geo- 
logical changes which they have undergone. With each succeeding 
stage the percentage of carbon increases and the oxygen content de- 
creases. The chemical changes are approximately as given in Table 1. 



TABLE L 

PROGRESSIVE CHANGE FROM PURE CELLULOSE TO ANTHRACITE. 



Substance. 



Pure cellulose 

Wood 

Peat 

Lignite 

Brown coal 

Bituminous coal 

Semi-bituminous coal 

Anthracite 

Graphite ..." 



Carbon. 


Hydrogen. 


Per Cent. 


Per Cent. 


44.44 


6.17 


52.65 


5.25 


59.57 


5.96 


66.04 


5.27 


73.18 


5.58 


75.06 


5.84 


89.29 


5.05 


91.58 


3.96 


100.00 





Oxygen. 



Per Cent. 
49.39 

42.10 
34.47 
28.69 
21.14 
19.10 
6.66 
4.46 



Of all the various grades of solid fuels coal is by far the most important. 
Origin of Coal : Bulletin No. 491, U. S. Geological Survey, p. 705. 

9. Composition of Coal. — All coals when separated into their ulti- 
mate chemical constituents are composed principally of varying pro- 
portions of carbon, hydrogen, oxygen, sulphur and refractory earths. 
Carbon and hydrogen are the only desirable elements from a combustion 
standpoint and the others may be considered impurities. The various 
combinations into which the carbon, hydrogen and oxygen are united 
are extremely complex and greatly influence the physical characteristics 
of the fuel. All of the carbon and hydrogen is not available for com- 
bustion since part of the carbon may be present as a carbonate and 
part of the hydrogen as water. A knowledge of the physical and chemi- 
cal characteristics of a fuel as determined in the laboratory is of great 
importance since it enables the engineer to determine in advance the 
fuels best suited for a given or proposed equipment. 

Proximate Analysis. — This analysis enables the engineer to pre- 
dict to a certain extent the behavior of the fuel in the furnace by giving 
the percentages of moisture, ash, fixed carbon and volatile matter. 

Moisture as obtained from this analysis is purely an arbitrary quantity 
based upon the loss in weight of a sample when maintained for approxi- 
mately one hour at a temperature of 220 deg. fahr. The material 
driven off in this manner is not all water since some of the volatile 
combustible may distill off; furthermore, all of the water may not be 



24 STEAM POWER PLANT ENGINEERING 

evaporated by this treatment. It is intended and does bring the ma- 
terial to a condition which can be dupUcated closely and represents 
a fixed basis for comparison. Moisture not only increases the cost of 
transporting and handling the fuel but also absorbs heat in the furnace 
which might otherwise be available for generating steam. Coal free 
from "moisture" is known as "dry coal.^^ '^ 

The residue which remains after the coal has been completely burned 
is classified as ash. It is derived from the inorganic matter in the 
coal, such as sand, clay, shale, "slate" and iron pyrites, and is composed 
largely of compounds of sihca, alumina, iron and lime, together with 
small quantities of magnesia. A large percentage of ash is undesirable 
since it reduces the heat value of the fuel, increases the cost of trans- 
portation and handling, necessitates disposal of refuse and often 
produces troublesome clinker. Coal free from moisture and ash is 
commonly designated as combustible though the nitrogen and oxygen 
included are not combustible. 

That portion of the carbon combined with hydrogen, and other 
gaseous compounds which are driven off the dry coal by the application 
of heat, constitutes the volatile combustible matter, or simply volatile 
matter. The term "volatile combustible" is a misnomer since a con- 
siderable fraction of the distilled gases consists of water vapor, carbon 
dioxide, nitrogen and other inert, non-combustible dilutants. The per- 
tinence of the "volatile matter" to the engineeer is obvious, since a 
high percentage indicates that special care must be observed in effect- 
ing smokeless combustion. 

The uncombined carbon or that portion which remains after the vola- 
tile matter has been driven off is known as fixed carbon. Fixed carbon, 
however, is not pure carbon since the carbonized residue contains in 
addition to the ash forming constituents, small amounts of hydrogen, 
oxygen, nitrogen and approximately half the original sulphur content. 
"Fixed carbon" is a measure of the relative coking properties of coals 
though in the commercial manufacture of coke or gas the yield of coke 
is several per cent higher than that obtained in the laboratory. In 

* "Moisture" as determined from the proximate analysis must not be confused 
with "air-drying loss." The primary purpose of air-drying is to reduce the mois- 
ture content to such a condition that there will not be rapid changes in the weight 
of the sample during the course of analysis; it simply shows the amount of moisture 
removed in order to bring the sample to a condition of equilibrium with respect to 
the moisture in the air of the room. "Air-drying loss" is the amount of moisture 
driven off when the sample, as received, is subjected to a temperature of 86 to 95 
deg. fahr. The drying process is continued until the loss in weight between two 
successive weighings made 6 to 12 hours apart does not exceed 0.2 per cent. See 
"Analysis of Coal in the United States, " Bulletin 22, 1913, Bureau of Mines. 



FUELS AND COMBUSTION 25 

the proximate analysis of coal the sulphur is included in the volatile 
matter, fixed carbon and ash. Sulphur occurs in coal as pyrites, sul- 
phate of iron, lime, and alumina, and in combination with the coal 
substance as organic compounds. Although classed as an impurity, 
sulphur has a heating value, when in the form of iron pyrites, of almost 
one-half that of the coal it replaces. For steaming purposes sulphur is 
objectionable only when its presence produces a badly clinkering ash. 

Proximate Analysis: Journal of the American Chemical Society, Vol. 21, p. 116; 
U. S. Bureau of Mines, Bulletins No. 22, 1913, and No. 85, 1914. 

Ultimate Analysis. — In the ultimate analysis the composition of 
the fuel is expressed in terms of its elementary constituents of carbon, 
hydrogen, oxygen, nitrogen and sulphur, and ash. The ultimate anal- 
ysis is of considerable importance in determining the more important 
heat losses incident to combustion, but an accurate analysis requires 
considerable time for its consummation and necessitates the services 
of a competent chemist. For that matter an accurate proximate 
analysis requires even more skill than the ultimate analysis, since 
the determination of hydrogen, carbon and nitrogen is not subject to 
the arbitrary conditions that must be maintained in the proximate 
analysis. But as ordinarily made the latter requires little apparatus 
and is within the skill of the average engineer. 

Both the ultimate and the proximate analyses may be expressed in 
terms of 

(1) ''Coal as received" or coal as fired, 

(2) "Coal, moisture free" or dry coal. 

(3) "Coal, moisture and ash free" or comhustible, 

(4) "Coal, moisture, ash and sulphur free." 

In the various fuel publications issued by the Bureau of Mines and 
the U. S. Geological Survey, the quoted terms are used almost ex- 
clusively, whereas in the Boiler Code advocated by the American 
Society of Mechanical Engineers and in most engineering literature 
the italicised terms are given preference. Engineers prefer to have the 
results based on coal as fired, since this represents the condition of 
the fuel as fed to the furnace. For convenience in comparing analyses 
the results are usually based on dry coal and combustible, but occasion- 
ally, as will be shown later, the "coal, moisture, ash and sulphur free" 
basis is of service. Analyses are readily converted from one basis to 
another as will be seen from the following example. 

Example 1. Given the proximate and ultimate analyses of a sample 
of coal as received. Transfer these analyses to the "moisture free" 
and "moisture and ash free" basis. Also transfer the ultimate analy- 
sis as received to the "moisture, ash and sulphur free" basis. 



26 



STEAM POWER PLANT ENGINEERING 



ILLINOIS COAL. 
(Carterville District. ) 





Coal as Re- 
ceived or Coal 
as Fired. 


Coal, Moisture 

Free or Dry 

Coal. 


Coal, Moisture 
and Ash Free, 
or Combus- 
tible. 




A. 


B. 


C. 


Fixed carbon 


50.19 

31.44 

10.61 

7.76 


54.42 
34.08 
11.50 


61.49 


Volatile matter 


38.51 


Ash 




Moisture ... 












100.00 


100.00 


100.00 



Column B = column A 4- (1 — proportional weight of moisture) 

= column A -i- 0.9224. 
Column C = column A -^ [1 — (proportional weight of moisture + 
ash)] 

= column A -^ 0.8163. 
For the ultimate analysis: 



Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Sulphur 

Ash 

Free moisture 



Coal as Received. 



A. 



66.55 
5.14 
1.32 

14.41 
1.97 

10.61 



100.00 



66.55 
4.28 
1.32 
7.51 
1.97 
10.61 
*7.76 



100.00 



Coal, 

Moisture 

Free. 



72.15 

4.64 
1.43 
8.14 
2.14 
11.50 



100.00 



Coal, 
Moisture 
and Ash 

Free. 



81.52 
5.24 
1.62 
9.21 
2.41 



100.00 



Coal, 
Moisture, 
Ash and 
Sulphur 

Free. 



83.54 
5.37 
1.66 
9.43 



100.00 



* From the proximate analysis. 



In the ultimate analysis of the coal as received (Column A) the free 
moisture of "Moisture" is included in the hydrogen and oxygen. Since 
the water is composed of one part hydrogen and eight parts oxygen, 
one-ninth of the moisture should be subtracted from the hydrogen 
and eight-ninths from the oxygen in order to include free moisture 
as a separate item, thus: 



Hydrogen (column Ai) = hydrogen (column A) 

ture 
= 5.14 - i X 7.76 

= 4.28. 



^ X per cent mois- 



FUELS AND COMBUSTION 27 

Oxygen (column Ai) = oxygen (column A) — .^ X per cent mois- 
ture 
= 14.41 - I X 7.76 
= 7.51. 
Column B = column Ai -^ (1 — proportional weight of moisture) 

= column Ai -f- 0.9224. 
Column C = column Ai -^ [1 — proportional weight of (moisture + 
ash)] 
= column Ai ^ 0.8163. 
Column D = column Ai -^ [1 — proportional weight of (moisture + 
ash + sulphur)] 
= column Ai ^ 0.7966. 

The term free hydrogen or available hydrogen is based on the assump- 
tion that all of the oxygen in the coal is combined with hydrogen in the 
proper ratio to form water, or, 

Oxygen O 

Free hydrogen = Total hydrogen ^^=^ — = H — - • 

o o 

All of the oxygen + -5- is the weight of the combined moisture, and the 

o 

sum of the free moisture and combined moisture is designated as the 
total moisture. 

Example 2. Determine the free hydrogen, combined moisture and 
total moisture for coal as fired, the analysis of which is given in Example 1. 

7 51 
Free hydrogen = 4.28 '-^ 

= 3.34. 

7 51 
Combined moisture = 7.51 H ^— 

o 
= 8.45. 
Total moisture = 7.76 + 8.45 
= 16.21. 

For most engineering purposes extreme accuracy is not necessary in 
determining the ultimate analysis, since the average commercial heat 
balance is in itself only approximate at the best, consequently recourse 
may be had to empirical formulas for approximating the weight of the 
chemical constituents from the proximate analysis, thus:* 

For hydrogen, H = V (y^^ - O.OI3), (1) 

in which 

H = the per cent of hydrogen in the combustible. 

V = the per cent of volatile matter in the combustible. 

* "Experimental Engineering," Carpenter & Diederichs, 1915, p. 507. 



28 



STEAM POWER PLANT ENGINEERING 



(2) 



(3) 



For nitrogen, 

N = 0.07 V for anthracite and semi-anthracite 
= 2.10 — 0.012 V for bituminous and hgnite. 

For total carbon (fixed ca^rbon + volatile carbon), 
C = F + 0.02 V2 for anthracite 

= F + 0.9 (V - 10) for semi-anthracite 
= F + 0.9 (V — 14) for bituminous coals 
= F + 0.9 (V - 18) for Ugnites 

in which 

C = per cent of total carbon in the combustible. 

F = per cent of fixed carbon as determined from the proximate 
analysis. 

V = as above. 

Sulphur in the coal increases the value of V, hence the calculated 
value of C is too high by practically the sulphur content of the com- 
bustible. 

Example 3. Calculate the ultimate analysis from the proximate 
analysis of the coal given in Example 1. 

.35 



H = 38.51 



V38.51 



')- 



0.013 ) = 5.33 per cent. (Analysis gives 



+ 10 
H = 5.24.) 
N = 2.10 - 0.012 X 38.51 

= 1.64 per cent. (Analysis gives N = 1.62.) 
C = 61.49 + 0.9 (38.51 - 14) 

= 83.55 per cent. (Analysis gives C = 81.52.) 

The ultimate analysis of the coal as received, neglecting the sulphur, is: 



H = 5.33 

N = 1.64 ^X 0.8163. 

C =83.55 , 

Ash (by analysis) ..... 

Moisture (by analysis) 

O (by difference) 



100.00 



Calculated Values, 


Actual Values, 


Per Cent. 


Per Cent. 


f 4.35 


4.28 


\ 1.33 


1.32 


168.20 


68.52* 


10.61 


10.61 


7.76 


7.76 


7.75 


7.51 



100.00 



Carbon + sulphur = 66.55 + 1.97 



.52. 



It will be seen that the agreement is fairly close with the exception 
of that for total carbon. As previously stated this is largely due to the 
fact that the sulphur content is practically all added to the total carbon. 
If the sulphur content of the coal is known, as in this case (2.41 per cent), 
correction can be made so that the final computed value for the total 
carbon is 83.55 — 2.41 = 81.14 per cent per lb. of combustible. 



FUELS AND COMBUSTION 



29 



This method of calculating the ultimate from the proximate analysis 
gives fairly accurate results for most coals but with some gracles of 
bituminous coals the results for H and C may be in error as much as 
5 per cent for each constituent. 

The average plant is not equipped with the necessary apparatus for 
making the proximate analysis, not alone the ultimate analysis, so that 
the preceding calculations are of Httle value to the engineer in charge. 
The proximate analysis is too cumbersome even for the large plant when 
a number of heat balances are required in a short time, as when trying 
out new fuels. In such cases the following method enables the engineer 
to approximate the ultimate analysis with sufficient accuracy for most 
practical purposes, provided the source of coal supply is known:* 

Bulletins Nos. 22 and 85 issued by the Bureau of Mines, contain a 
large number of ultimate analyses of coals from all parts of the country. 
A study of the data will show that coals from any given locality have 
practically the same analysis when expressed on a "free from moisture, 
ash and sulphur ^^ basis; hence, it is principally a question of deter- 
mining the amount of free moisture and ash in the sample (a compara- 
tively simple test) and in assuming the sulphur content. Since the 
percentage of sulphur is not uniform some error may be introduced in 
making this assumption but it is negUgible as far as the average com- 
mercial heat balance is concerned. This method of obtaining the 
ultimate analysis is best illustrated by an example. 

Example 4. Assume that a sample of Illinois coal (analysis as per 
Example 1) is available and that the ash and moisture determinations 
only have been made. Approximate the ultimate analysis from the 
average "moisture, ash and sulphur free" analysis of Illinois coals. 

The average of a number of Illinois coals j as recorded in the Govern- 
ment bulletin referred to is: 



Combined 
Moisture. 


Free Hydrogen. 


Carbon. 


Nitrogen. 


11.94 


4.14 


82.4 


1.52 



Assuming the per cent of sulphur in the coal under consideration to be 
the average of Illinois coals as recorded in the Government bulletins 
(S = 2.84 per cent), the total free moisture, ash and sulphur would be 
7.76 + 10.61 + 2.84 = 21.2 per cent; and the ''free from moisture, 
ash and sulphur" content, 100 - 21.2 = 78.8 per cent. The ultimate 
analysis of the coal as received may then be calculated as follows: 

* P. W. Evans, Armour Engineer, May, 1915, p. 301. 
t Moisture, ash and sulphur free. 



30 



STEAM POWER PLANT ENGINEERING 





Calculated Values, 
Per Cent. 


Actual Values, 
Per Cent. 


Combined moisture, 11.94 X 0.788 

Free moisture (by test) 


9.40 

7.76 

3.26 

64.98 

1.19 

10.61 

*2.80 


8.45 
7.76 


Free hydrogen 4 14 X 788 


3.34 


Total carbon, 82 3 X . 788 


66.55 


Nitrogen, 1.52 X 0.788 


1.32 


Ash (bv test) 


10.61 


Sulphur 


1.97 




100.00 


100.00 



* By assumption. 

The agreement between calculated and actual values for most Illinois 
coals is much closer than in this particular example. The splendid 
work of the U. S. Bureau of Mines will soon place at the disposal of the 
public complete analyses of all the coal fields in the country and the 
error in assuming the average values of an entire state, as in the preced- 
ing example, may be greatly reduced by taking the average values for 
the particular field in which the coal under consideration is mined. 

Ultimate Analysis: See references under " Approximate Analysis." 

Methods of Determining Sulphur Content of Fuels: U. S. Bureau of Mines, Tech- 
nical Paper 26, 1912. 

Methods of Analyzing Coal and Coke: U. S. Bureau of Mines, Technical Paper 8, 
1913. 

The Coking of Coal at Low Temperatures: University of Illinois Bulletin, Vol. XII, 
No. 39, 1915. 

The Analysis of Coal with Phenol as a Solvent: University of Illinois Bulletin, 
No. 76, 1915. 



10. Classification of Coals. — Coals and allied substances have been 
variously classified according to 

1. Oxygen-hydrogen ratio, or Gruner's classification. 

2. Fixed carbon and volatile combustible matter. 

3. Fuel ratio, or the ratio of the fixed carbon to the volatile combus- 
tible matter. 

4. Calorific value. 

5. Fixed carbon. 

6. Total carbon. 

7. Hydrogen. 

8. Carbon-hydrogen ratio, or the ratio of the total carbon to the 
hydrogen. 



FUELS AND COMBUSTIOM 
Gruner's classification is as follows: 

(Eng. and Min. Jour., July 25, 1874.) 



31 



Anthracite. 
Bituminous 
Lignite. . . . 



Ratio |. 




Ratio g. 


1 to 0.75 
4tol 
5 


Peat 

Wood 

Cellulose 


6 to 5 

7 
8 



Kent's classification, according to the constituents of the combustible, 
is as follows (Steam Boiler Practice) : 



Anthracite. 

Semi-anthracite 

Semi-bituminous 

Bituminous — Eastern , 
Bituminous — Western 
Lignite 



Per Cent of Dry Combustible. 



Fixed Carbon. 



97 to 92.5 
92.5 to 87.5 
87.5 to 75 
75 to 60 
65 to 50 
Under 50 



Volatile Matter. 



3 to 7.5 

7.5 to 12.5 

12.5 to 25 

25 to 40 

35 to 50 

Over 50 



Gruner's, Kent's, and other schemes of classification outlined above, 
with the exception of the carbon-hydrogen ratio, are more or less 
unsatisfactory, since the groups are not as clearly defined as indicated 
and overlap to a considerable extent. 

The U. S. Geological Survey proposes the following classification 
according to the carbon-hydrogen ratio which appears to apply satis- 
factorily to all grades of coal. 



(Compiled from Report of Government Coal Testing Plant, Professional Paper No. 48, 1906.) 



Groifp. 


Class. 


Example. 


Carbon-hydrogen 
Ratio. 


A 


Graphite . 






B 


Anthracite 

Anthracite. . . 


*Buck Mountain, Pa 


to 30 


c 


*Scranton, Pa. 


30 to 26 


D 


Semi-anthracite — 
Semi-bituminous . . . 

Bituminous 

.do 


*Bernice Basin, Pa. . . 


26 to 23 


E 


Spadra Bed, Ark. . 


23 to 20 


F 

G 


New River, W. Va 

Connelsville Field, Pa 

Marion County, III 


20 to 17 
17 to 14 4 


H 


...do 


14 4 to 12 5 


I 


...do 


Red Lodge, Mont 


12 5 to 11 2 


J 


Lignite. 


Gallup Field, N. M 


11 2 to 9 3 


K. 


Peat . . . 




9 3 to 


L 


Wood 




7 2 











Not included in Government's Report. 



32 



STEAM POWER PLANT ENGINEERING 



In its various bulletins the U. S. Bureau of Mines uses the following 
arbitrary classification which is virtually based on the fuel ratio: 

Anthracite Bituminous 

Semi-anthracite Sub-bituminous 

Semi-bituminous Lignite 

Classification of Coals: U. S. Geographical Survey Bulletin 541, 1914; Prac. Engr. 
U. S., Jan., 1910: Mines and Minerals, Feb., 1911. 

11. Anthracites. — These are the best coals and consist almost en- 
tirely of carbon; they contain very little hydrocarbon and burn with 
Uttle or no smoke, are slow to ignite, burn slowly, and break into small 
pieces when rapidly heated. They require a very large grate of about 
twice the surface necessary for bituminous coal. Large sizes may be 
burned in almost any kind of a furnace and with moderate draft. 

TABLE 2. 

COMPOSITION OF TYPICAL AMERICAN ANTHRACITE COALS. 



Proximate analysis: 

Water 

Volatile matter . . . 

Fixed carbon 

Ash 



Ultimate analysis: 

Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Sulphur 

Ash 



Calorific value: 
Calorimeter .... 
Dulong's formula. 



Classification: 
Carbon-hydrogen ratio . 
Fuel ratio 



£ 



0.84 

6.67 

85.66 

6.83 



100.00 

90.66 
1.73 



0.78 
6^83 



100.00 

13,980 
14,194 



52.5 
12.9 



%i 



3.45 

2.75 

87.90 

5.90 



100.00 



2.04 
0.90 
1.95 
0.35 
5.90 



100.00 

13,950 
14,103 



42.5 
32.0 



t 

1.37 

3.59 

89.11 

5.93 



100.00 

87.70 
2.56 
1.03 
2.26 
0.56 
5.89 



100.00 



14,217 



34.4 
29.9 



T 



t 

1.97 

4.35 

86.49 

7.19 



100.00 

85.66 

2.78 
0.77 
2.87 
0.64 

7.28 



100.00 



14,038 



30.9 
11.0 






t 

2.08 

7.27 

74.32 

16.33 



100.00 

75.21 

2.81 
0.80 
4.08 
0.77 
16.33 



100.00 

12,472 
12,426 



26.7 
10.2 



t 
1.50 

7.84 

81.07 

9.59 



100.00 

83.20 
3.29 
0.95 
2.45 
0.50 
9.61 



100.00 



14,003 



25. 
10.4 



Authority not stated. f H. J. Williams. t U. S. Geological Survey. 



For smaller sizes a thinner bed has to be carried unless a strong draft 
is used. There is difficulty in keeping a thin bed free from air-holes. 



FUELS AND COMBUSTION 



33 



When possible the coal should be at least six inches deep on the grate. 
On account of the large percentage of ash in the smaller sizes, the fire 
requires frequent cleaning. Anthracites do not require ''shcing" and 
should be disturbed only when cleaning is necessary. Nearly all an- 
thracites, with some unimportant exceptions, come from three small 
fields in eastern Pennsylvania. On account of the limited supply and 
the great demand for domestic purposes, sizes over ''pea coal" are 
prohibitive in price for steam power plant use. Table 2 gives the 
composition and classification of a number of typical American anthra- 
cite coals, and Table 3, one of the standard divisions of mesh according 
to which they are classed and marketed. Specific gravity, 1.4 to 1.6; 
fuel ratio, not less than 10. 

Burning No. S Buckwheat: Power, Dec. 27, 1901; Mar. 21, 1911. Burning 
Anthracite Culm of Poor Quality: Trans. A.S.M.E., 7-390. Anthracite Culm Bri- 
quets, Am. Inst. Min. Engrs., Bulletin, Sept., 1911. Calorific Value of Anthracite: 
Mines and Minerals, Sept., 1911, Preparation of Anthracite: Am, Inst, Min. 
Engrs,, Bulletin, Oct,, 1911. Stoking Small Anthracite Coal: Power, Oct. 19, 1916, 
p,540. 

TABLE 3. 

SIZES OF ANTHRACITE COAL. 
A.S.M.E, Code of 1915, 





Size, 


Diameter of Opening Through or Over 
which Coal Will Pass, Inches. 




Through, 


Over. 


Broken 


1 




31 


Egg 


2^ 


Stove 


If 


Chestnut . 




Pea 


%_ 


*No. 1 Buckwheat, , , 


^5^ 


*No. 2 Buckwheat. . , 


4 


*No. 3 Buckwheat 


Culm 


32 







* The terms " Buckwheat," "Rice," and "Barley," respectively, are used in some localities instead 
of No. 1, No. 2, and No. 3 Buckwheat. 



12. Semi-anthracites. — These coals kindle more readily and burn 
more rapidly than the anthracites. They require little attention, 
burn freely with a short flame and yield great heat with little clinker 
and ash. They are apt to split up on burning and waste somewhat in 
falUng through the grate. They swell considerably, but do not cake. 
They have less density, hardness and metallic luster than anthracite, 
and can generally be distinguished by their tendency to soil the hands, 



34 



STEAM POWER PLANT ENGINEERING 



while pure anthracite will not. Semi-anthracites are not of great im- 
portance in the steam power plant field on account of the limited sup- 
ply and high cost. They are found in a few small areas 'in the western 
part of the anthracite field. Specific gravity, 1.3 to 1.4. Fuel ratio 
6 to 10. 

13. Semi-bituminous. — These coals are similar in appearance to semi- 
anthracite, but they are somewhat softer and contain more volatile 
matter. They have a very high heating value, have a low moisture, 
ash and sulphur content, are readily burned without producing ob- 
jectionable smoke and rank among the best steaming coals in the 
world. The supply is limited and on account of high cost, except in 
the immediate vicinity of the mines, they are not generally used for 
power purposes. The centers of production are the Pocahontas and 
New River fields of Virginia and West Virginia, Georges Creek field of 
Maryland, Windber field of Pennsylvania and the western end of the 
Arkansas field. Table 4 gives the composition and classification of a 
number of typical American semi-anthracite and semi-bituminous coals. 
Fuel ratio 3 to 6 or 7. 



TABLE 4. 

COMPOSITION OF TYPICAL AMERICAN SEMI-ANTHRACITE AND SEMI-BITUMI- 

NOUS COALS. 



Proximate analysis: 

Water 

Volatile matter. . 

Fixed carbon 

Ash 



Ultimate analysis: 

Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Sulphur 

Ash 



Calorific value: 

Calorimeter 

Dulong's formula 

Classification: 

Carbon-hydrogen ratio] 

Fuel ratio 



wP-i 

I 



1.57 

9.40 

83.69 

5.34 

100.00 

85.46 
3.72 
1.12 
3.45 
0.91 
5.34 



100.00 



14,552 

23.0 

8.5 



¥ 








* 

2.36 


t 

4.07 


1 

1.42 


t 
1.53 


12.68 


16.34 


20.72 


21.54 


72.88 


68.47 


70.05 


71.88 


12.08 


11.12 


7.81 


5.05 


100.00 


100.00 


100.00 


100.00 


76.44 


76.51 


81.95 


82.87 


3.82 


4.27 


4.30 


4.76 


1.37 


1.00 


1.29 


1.68 


4.30 


6.59 


3.68 


4.99 


1.99 


0.51 


0.97 


0.65 


12.08 


11.12 


7.81 


5.05 


100.00 


100.00 


100.00 


100.00 


13,259 
13,273 


13,509 
13,329 


14,686 
14,363 


14,807 
14,691 


20.7 


19.6 


19.0 


17.8 


5.7 


4.2 


3.4 


3.3 



0.44 
18.76 
73.15 

7.65 



100.00 

80.32 
4.88 
1.46 
4.69 
1.00 
7.65 



100.00 



14,432 

16.5 
3.9 



• Authority not stated. t U. S. Geological Survey. t W. Va. Geological Survey. 
§ H. J. Williams. || Based on air-dried sample. 



FUELS AND COMBUSTION 35 

14. Bituminous. — These coals are the most widely distributed and 
the most extensively used fuel in steam power plant engineering. They 
contain a large and varying amount of volatile matter and burn freely 
with the production of considerable smoke unless carefully fired. Their 
physical properties vary widely and they are commonly classified as 

1. Dry, or free-burning bituminous. 

2. Bituminous caking. 

3. Long-flaming bituminous. 

1. Dry bituminous coals are the best of the bituminous variety for 
steaming purposes. They are hard and dense, black in color, but some- 
what brittle and splintery. They ignite readily, burn freely with a 
short clean bluish flame and without caking. Specific gravity, L25 to 
L40. 

2. Bituminous caking coals swell up, become pasty and fuse together 
in burning. They contain less fixed carbon and more volatile matter 
than the free-burning grades. Caking coals are rich in hydrocarbon 
and are particularly adapted to gas making. The flame is of a yellowish 
color. Specific gravity, about L25. 

3. Long-flaming bituminous coals are similar in many respects to 
the caking coals but contain a larger percentage of volatile matter. 
They burn freely with a long yellowish flame. They may be either 
caking, non-caking or splintery. They are very valuable as a gas coal, 
and are little used for steaming purposes. Specific gravity, about L2. 

Table 5 gives the composition and classification of a number of typical 
American bituminous coals. 

For sizes of bituminous coal see paragraph 38. 

Mineral Resources of the United States: U. S. Geological Survey, 19 IL 

Analyses of Coals: Bui. No. 22, U. S. Bureau of Mines, 1913. 

Analyses of Mine and Car Samples: Bui. No. 85, U. S. Bureau of Mines, 1914. 

Report of the United States Fuel-Testing Plant at St. Louis, Mo.: Bui. No. 332, 
U. S. Geological Survey, 1908. 

Index of Mining Engineering Literature: W. R. Crane, John Wiley & Sons. 

Coal Mines of the United States: Peabody Atlas, A. Bement, Chicago, 111. 

Coking and Caking Coal: Power, March 28, 1916, p. 432. 

Dry Preparation of Bituminous Coal at Illinois Mines: Univ. 111. Bui. No. 43, 
June 26, 1916. 

Fuels for Steam Boilers: Power, Mar. 28, 1916, p. 454. 

15. Sub-bituminous Coals. — The term sub-bituminous has been 
adopted by the U. S. Geological Survey and the Bureau of Mines for 
what has generally been called ''black hgnite. " These coals are not Hg- 
nitic in the sense of being woody and many of them approach the lowest 
grade of bituminous coals for fuel purposes. It is difficult to separate 



36 



STEAM POWER PLANT ENGINEERING 



tf 
w S 



•BMOJ 



'BuquiBQ 



HI 

'U88J50Q 



'UIH qaiH 



•BAVOJ 
'ajBp^pp'B'^ 



•SBSn'BX 



8 



8 8! 



O (NCO 

o 



oo ^ 



8 



o s 



O CO CO' 

OQO 






ooooai"^ o 



8 



•3{8aj09SJOJJ 






•pni 



8 



8 



(MOO Tt^ • 

ooo 









iO(M 

00 Oi 



o c<r<M 



O 00 lO 
O ^ lO 

r-(05 



o ^^ 



•qoiH 



'AajiBA 
3uiJ{0ojj 



8 



8 



O lO'* 
Ort< 

8 CO<N 



'^jodaajj 



O 0(M 
OOi 






"1^ ri 



to 00 

Tl^ r-l 



s; 



§ 32 3' 



(M(N ^ 



§ 81 



•^+H •' 



iOt-i 



'^co ;:; 



oco 

0(N 



«. r^ hC IJ rn" »H 

.5 '^ ^'JZ X 3 to 



03 
o J^ ^ to w 



o 

O 133 



FUELS AND COMBUSTION 



37 



this class from bituminous coals and lignites by any of the classifications 
outhned at the beginning of paragraph 10. They are not woody in 
texture and are black in color, which enables them to be readily dis- 
tinguished from the lignites. When exposed to the weather they slack 
considerably, a feature which distinguishes them from the bituminous 
coals. Sub-bituminous coals are found in most of the western fields. 
16. Lignite, or Brown coal, is a substance of more recent geological 
formation than coal and represents a stage in development intermediate 
between coal and peat. Its specific gravity is low, 1.2, and when freshly 
mined contains as high as 50 per cent of moisture. It is non-caking, and 
on exposure to air, slackens or crumbles. The lumps check and fall 
into small irregular pieces with a tendency to separate into extremely 
thin plates. It deteriorates greatly during storage or long transpor- 
tation. Lignite, as mined, is a low-grade fuel with a calorific value of 
about one-half that of good coal. When properly prepared and com- 
pressed into briquettes, lignite becomes an excellent fuel, resists weather- 
ing satisfactorily, permits handUng and transportation without ex- 
cessive deterioration and is practically smokeless. The superiority of 
briquettes over raw lignite is shown in Table 6: 

TABLE 6. 

IMPROVEMENT OF HEAT VALUE BY BRIQUETTING.* 





Moisture 


Heat ^'alue per Pound. 


Source 


In Raw 
Lignite. 


In Bri- 
quettes. 


Removed. 


Raw- 
Lignite. 


Briquettes. 


Increase. 


Texas 


Per Cent. 
33.0 
40.0 
42.0 
40.0 


Per Cent. 
9.0 
12.0 
10.0 

10.0 


Per Cent. 

24.0 
28.0 
32.0 
30.0 


B.t.u. 

6840 
6241 
6079 
6080 


B.t.u. 

9336 
9354 
9355 
9264 


Per Cent. 
36 5 


North Dakota 

North Dakota 

California 


50.0 

54.0 
52.4 



* Bulletin No. 14, U. S. Bureau of Mines, p. 48. 

The most extensive Ugnite deposits are situated long distances from 
fields of high-grade coal, and their use is at present Umited to these 
regions. 

North Dakota Lignite as a Fuel for Power Plant Boilers: Bui. No. 2, 1910, U. S. 
Bureau of Mines. Briquetting Tests of Lignite: Bui. No. 14, 1911, U. S. Bureau of 
Mines. General data pertaining to lignite fuels, Engr. U. S., Jan., 1910. 



17. Peat, or Turf, is formed by the slow carbonization under water 
of a variety of accumulated vegetable materials. It is unsuitable for 
fuel until dried. Peat, as ordinarily cut and dried, is too bulky for 



38 



STEAM POWER PLANT ENGINEERING 



commercial competition with coal, and is used only where coal is pro- 
hibitive in price. When properly prepared and compressed into bri- 
quettes peat is an excellent fuel. In Russia, Germany, and Holland 
peat briquettes have passed the experimental stage and several millions 
of pounds are manufactured annually. Peat is used but little in this 
country at present, though the deposits are extensive and widely dis- 
tributed, but its possibiUties are beginning to attract the attention of 
engineers. The proportion in which the various primary constituents 
exist in dried peat is approximately as follows: Per Cent. 

Fixed carbon 35 

Volatile matter 60 

Ash 5 

Peat: Prac. Engr. U. S., Jan., 1910, p. 21; Bui. No. 16, U. S. Bureau of Mines, 
1911; Power, Sept. 6, 1910; Eng. and Min. Jour., Nov. 22, 1902; Feb. 7, 1903, 
Jour. Am. Peat See, July, 1911; Elec. Rev., Mar. 22, 1912; Min. and Eng. Wld., 
Nov. 28, 1911. 

TABLE 7. 

COMPOSITION OF TYPICAL AMERICAN SUB-BITUMINOUS COALS AND LIGNITES.* 

(Run of Mine.) 



S § 



Proximate analysis: 

Water 

Volatile matter. . 

Fixed carbon 

Ash 



Ultimate analysis: 

Hydrogen 

Carbon 

Nitrogen 

Oxygen 

Sulphur 

Ash 



Calorific value: 

Calorimeter 

Dulong's formula 

Classification: 

Carbon-hydrogen ratiof 

Fuel ratio 





8^ 


^^ 


. 


.—- 




III 




11 


III 


11.05 


12.29 


33.71 


18.68 


36.78 


35.90 


34.58 


29.25 


34.88 


28.16 


42.08 


46.14 


29.76 


40.45 


29.97 


10.97 


6.99 


7.28 


5.99 


5.09 


100.00 


100.00 


100.00 


100.00 


100.00 


5.37 


5.82 


6.79 


6.07 


6.93 


59.08 


63.31 


45.52 


57.46 


41.87 


1.33 


1.03 


0.79 


1.15 


0.69 


21.52 


22.22 


42.09 


28.78 


44.94 


1.73 


0.63 


0.53 


0.55 


0.48 


10.97 


6.99 


7.28 


5.99 


5.09 


100.00 


100.00 


100.00 


100.00 


100.00 


10,539 


11,252 


7348 


10,143 


7002 


10,355 


11,153 


7177 


9,948 


6944 


11.50 


11.20 


10.90 


9.80 


9.60 


1.17 


1.09 


1.02 


1.16 


1.06 



22.63 

35.68 

37.19 

4.50 

100.00 

6.39 
54.91 

1.02 
32.59 

0.59 

4.50 

100.00 

9734 
9478 

9.40 
1.05 



* Compiled from Government Report, U. S. Geological Survey, 
t Based on air-dried analysis. 

18. Wood, straw, Sawdust, Bagasse, Tanbark. — In certain localities 
cordwood is still used as a fuel, but the steadily increasing values of 
even the poorest qualities are rapidly prohibiting its use for steam- 



FUELS AND COMBUSTION 



39 



generating purposes. Sawdust, shavings, tanbark and other waste 
products of wood are burned under boilers in situations where such 
disposition nets the best financial returns. Recent progress, however, 
in industrial chemistry shows that ethyl and wood alcohols and other 
valuable by-products can be cheaply made from sawdust, shavings, 
slashings and similar waste material, and it is not unUkely that their 
use for steaming purposes will be unheard of in a comparatively few 
years. Table 8 gives the physical and chemical characteristics of a 
number of woods. 

TABLE 8. 

PHYSICAL AND CHEMICAL PROPERTIES OF WOODS, STRAW AND TANBARK. 
(Prac. Engr. U. S., Jan., 1910.) 







r6 

8 

W) 

1 


Equivalent Weight 
of Coal. 13.500 
B.T.U. 


6 

IS 
6 


1 


P 

s 




6 


Calorific Value, 
-(• B.T.U. per 
Pound. 


< 


Ash 


46 
43 
45 
42 
41 
35 
25 
53 
49 
59 
52 
45 
25 
36 
36 
25 
35 
25 


3520 
3250 
2880 
3140 
2350 
2350 
1220 
4500 
3310 
3850 
3850 
3310 
1920 
2130 
2130 
1920 
3310 
1920 


1420 

1300 

1190 

1260 

940 

940 

580 

1800 

1340 

1560 

1540 

1340 

970 

1050 

1050 

970 

1340 

970 












5450 
5400 
5580 
5420 
5400 
5400 
6410 
5400 
.5460 


Hutton 


Beech 

Birch 

Cherry 


49.36 
50.20 


6.01 
6.20 


42.69 
41.62 


0.91 
1.15 


1.06 
0.81 


Sharpless 
Hutton 












Sharpless 
<< 


Elm 


























Hickory 

Maple, Hard 
Oak Ijive 












Sharpless 
Hutton 






















5460 
5400 
5460 
6830 
6660 


<( 


'' White. 
" Red 


49.64 


5.92 


41.16 


1.29 


1.97 


Rankine 
Hutton 


Pine, White 
" Yellow 












u 












ft 


Poplar 

Spruce 

Walnut 


49.37 


6.21 


41.60 


0.96 


1.86 


6660 
6830 
5460 
6830 


(I 












<( 


Willow .... 


49.96 


5.96 
6.06 


39.56 


0.96 
1.05 


3.37 
1.80 


Rankine 


Average. . 


49.70 


41.30 














Straw: 
Wheat . . . 
Barley . . . 


•X- 

00 

s 

CO 




Water 
16.00 
15.50 


35.86 
36.27 


5.01 
5.07 
5.04 


37.68 
38.26 
37.97 


0.45 
0.40 
0.42 


5.00 
4.50 
4.75 


5155 


Clark 


Average 


15.75 


36.06 




Tanbark 

Dry 








51.80 


6.04 


40.74 




1.42 


6100 


Myers 









* Compressed. 



t Green Fuel. 



Wood as Fuel: Prac. Engr. U. S., Jan., 1910, p. 805; Power & Engr., June 30, 
1908, p. 1015; Power, Dec, 1908, p. 772. 

Burning Sawdust: Prac. Engr. U. S., Jan., 1910, p. 48; Power & Engr., April 7, 
1908, p. 536; Oct. 13, 1908, p. 613; Jour, of Elec, Oct., 1905. 



40 



STEAM POWER PLANT ENGINEERING 



TABLE 9. 

HEAT VALUES OF BAGASSE AND VARIATION WITH DEGREE OF EXTRACTION. 



II 


1 1 


Fiber. 


Sugar. 


Molasses. 


1^ 

& . 

Hi 


Heat Required to 

Evaporate the Water 

Present. B.T.U. 


1 


Lb. Bagasse Required 
to Equal lib. Coal of 
14,000 B.T.U. Cal- 
orific Power. 


Coal Equivalent per 

Ton of Cane. 

Pounds. 


2i 

E 








It 






c 


3" . 


1 


90 


0.00 


100.00 


8325 










8325 




8325 


1.68 


119 


2465° 


85 


28.33 


66.67 


5552 


3.33 


240 


1.67 


116 


5900 


339 


5561 


2.52 


119 


2236 


80 


42.50 


50.00 


4160 


5.00 


361 


2.50 


174 


4697 


509 


4188 


3.34 


120 


2023 


75 


51.00 


40.00 


3330 


6.00 


433 


3.00 


209 


3972 


611 


3361 


4.17 


120 


1862 


70 


56.67 


33.33 


2775 


6.67 


482 


3.33 


232 


3489 


679 


2810 


4.98 


120 


1732 


65 


60.71 


28.57 


2378 


7.15 


516 


3.57 


248 


3142 


727 


2415 


5.80 


121 


1612 


60 


63.75 


25.00 


2081 


7.50 


541 


3.75 


261 


2883 


764 


2119 


6.61 


121 


1513 


55 


66.12 


22.22 


1850 


7.78 


562 


3.88 


270 


2682 


792 


1890 


7.40 


121 


1427 


50 


68.00 


20.00 


1665 


8.00 


578 


4.00 


278 


2521 


815 


1706 


8.21 


122 


1350 


45 


69.55 


18.18 


1513 


8.18 


591 


4.09 


284 


2388 


833 


1555 


9.00 


122 


1284 


40 


70.83 


16.67 


1388 


8.33 


601 


4.17 


290 


2279 


849 


1430 


9.79 


123 


1222 


25 


73.67 


13.33 


1110 


8.67 


626 


4.33 


301 


2037 


883 


1154 


12.13 


124 


1077 


15 


75.00 


11.77 


980 


8.82 


637 


4.41 


307 


1924 


899 


1025 


13.66 


124 


1002 





76.50 


10.00 


832 


9.00 


650 


4.50 


313 


1795 


916 


879 


15.93 


126 


906 



Bagasse, or megass, is recuse sugar cane and is used as a fuel on the 
sugar plantations. Its tteat value depends upon the proportions of 
fiber, molasses, sugar and water left after the extraction. The heat 
furnished by the different constituents is about as follows: Fiber, 
8325 B.t.u. per pound; sugar, 7223 B.t.u. per pound; and molasses, 
6956 B.t.u. per pound. Table 9 gives the heat value of bagasse and 
variation with the degree of extraction. A typical furnace for burning 
bagasse is shown in Fig. 108. 

Bagasse as Fuel: Prac. Engr. U. S., Jan., 1910; Engng., Feb. 18, 1910. 
Bagasse Drying: E. W. Kerr, Louisiana Bui. No. 128, June, 1911. 

Tanbark is usually quite moist; the amount of moisture varies with 
the leaching process used and averages around 65 per cent. In this 
condition it has a heat value of about 4300 B.t.u. per pound. If per- 
fectly dry its heating power is approximately 6100 B.t.u. per pound. 
As in the case of all moist fuels, tanbark must be surrounded by heated 
surfaces of sufficient extent to insure drying out the fresh fuel as thrown 
on the fire. A, /Successful furnace for burning tanbark is shown in Fig. 
109. 

Tanbark as a Boiler Fuel: Jour. A.S.M.E., Feb., 1910, p. 181; Jour. A.S.M.E., 
Oct., 1909, p. 951; Prac. Engr. U.S., Jan., 1910. 
Burning Coke Breeze: Power, July 4, 1916, p. 2. 



FUELS AND COMBUSTION 41 

19. Combustion. — To the engineer combustion means the chemical 
union of the combustible of a fuel and the oxygen of the air at such a 
rate as to cause rapid increase in temperature. The depreciation in 
heat value of bituminous coal subjected to ''weathering" is due to com- 
bustion, but the rate at which the combustible unites with the oxygen 
is so slow that the heat is dissipated and there is practically no increase 
in temperature. When the combustible elements unite with oxygen 
they do so in definite proportions, which are always the same, and the 
union liberates a fixed quantity of heat independent of the time occupied. 
Theoretically combustion is a simple process as it is only necessary to 
bring each particle of fuel previously heated to the kindling tempera- 
ture in contact with the correct amount of oxygen and the combustion 
will be complete, the fuel oxidizing to the highest possible degree. In 
practice, however, the size and character of fuel, type of furnace, draft, 
impurities in the fuel, and the mechanical difficulties affect combus- 
tion to such an extent as to render oxidation more or less incomplete. 

When heat is applied to coal, combustion takes place in three separate 
and distinct stages: 

1. Absorption of heat. A fresh charge of fuel when thrown on a 
fire must first be brought to the kindling point in order that chemical 
action may take place. The temperatures necessary to cause this 
union of oxygen and fuel are approximately as follows: 

Deg. Fahr. Deg. Fahr. 

Lignite dust 300 Cokes 800 

Sulphur 470 Anthracite lump 750 

Dried peat 435 Carbon monoxide 1211 

Anthracite dust 570 Hydrogen 1 100 

Lump coal 600 

(Stromeyer, Marine Boiler Management and Construction, p. 93.) 

The amount of heat required to reahze the kindling temperature is 
greatly increased by the water content of the fuel since practically all of 
the free moisture must be evaporated before this temperature is reached. 

2. Vaporization of the hydrocarbon portion of the fuel and its com- 
bustion, the hydrocarbons consisting principally of ethylene gas, C2H4, 
methane gas, CH4, tar, pitch, naphtha and the Uke. As these gases are 
driven off they become mixed with the entering air, and the carbon 
and hydrogen unite with the oxygen, forming carbon dioxide, CO2, and 
water vapor, H2O, respectively, and give up heat in doing so. If 
volatile sulphur is present it unites with oxygen, forming sulphur 
dioxide, SO2, and also gives up heat. If insufficient oxygen is present 
for complete oxidation, the carbon may burn to carbon monoxide, CO, 
and only a small portion of the available heat be liberated. 



42 STEAM POWER PLANT ENGINEERING 

3. Combustion of the solid or carbonaceous portion of the fuel. 
After the hydrocarbons have been driven off and oxidized the remain- 
ing solid matter is composed chiefly of carbon and ash. The carbon 
unites with the oxygen, forming carbon dioxide, carbon monoxide, or 
both, depending upon the completeness of combustion. The ash, of 
course, remains unconsumed. 

In commercial practice the requirements for perfect combustion are 
a surplus of air, a thorough mixture of the fuel particles with the air, 
and a high temperature. The surplus of air above theoretical require- 
ments should be kept to a minimum, but even in the most scientifically 
designed furnace some excess is essential on account of the difficulty 
of properly mixing the gases, since the currents of combustible gases 
and air are apt to be more or less stratified. The products of combustion 
must be maintained at the kindling temperature until oxidation is com- 
ple'te, otherwise the carbon will be wasted as carbon monoxide or as 
smoke. The final products of combustion as exhausted by the chimney 
should consist only of carbon dioxide, water vapor, oxygen, nitrogen, 
and the oxides of impurities in the fuel. 

As previously stated when the combustible elements unite with oxy- 
gen they do so in definite proportions, called the combining weights, 
which are always the same, for a given reaction, and the union produces 
a fixed quantity of heat. Thus in the complete combustion of carbon, 
12 pounds of carbon unite with 32 pounds of oxygen, forming 44 pounds 
of carbon dioxide; hence, one pound of carbon will form 

C + O2 12 + 2 X 16 „2 A fnr^ 

— Y^ — - = -^ = 3f pounds of CO2 

and the heat of combustion will be about 14,540 B.t.u. per pound of 
carbon thus consumed. (The heat value for carbon appears to depend 
upon the method of preparation and ranges according to various author- 
ities from 14,220 to 14,647 B.t.u. per pound.) 

If combustion is incomplete and the carbon burns to carbon monoxide, 
one pound of carbon will form 

2 (^ -I- O 24 -I- "^2 
^"^ ' = 24: ^ ^^ pounds of CO and liberates 4380 B.t.u. 

Similarly, in burning to H2O one pound of hydrogen will form 
2H2 + O2 2 X 2 + 32 _ , „„^ 

2H. = 2X2 = ^ P^^^^' ^^ ^'^' 
(The exact figures, based upon the relative molecular weights, as adopted 
by the International Committee on Atomic Weights, are 
2 X 2.016 + 32 



2 X 2.016 



= 8.94 pounds. 



FUELS AND COMBUSTION 



43 



TABLE 10. 

DATA RELATIVE TO ELEMENTS MOST COMMOXLY MET WITH IX COXXECTIOX 
WITH COMBUSTIOX OF FUEL. 



Substance. 


Molecular 
Formula. 


Relative 

Molecular 

Weight, 

Oxygen 

=32. 


Chemical Reactions. 


Weight per 
Pound of Sub- 
stance in First 
Column. 




Oxygen 

3.08 

1^33 
2.66 

0^57 

8.0 

4.0 

3^43 

ro 


A.r. 


Acetylene 


C0H2 


26.02 

*12;o" 
*12.0 

44.0 

28.0 
2.016 

16.03 

28.02 

28.03 

32.0 
*32.07 

18.02 


2 C2H2+5 O2 = 4 CO2+2 H2O 


13.35 


Air 




Ash 








Carbon 


*c 

*c 
C02 

CO 

H2 
CH4 

N2 
C2H4 
02 
*s 
H20 


2C+02 = 2CO 
2C+2 02 = 2C02 


5.78 


Carbon .... 


11.58 


Carbon dioxide 




Carbon monoxide . 
Hydrogen 


2CO+02 = 2C02 
2H2 = 02 = 2H20 
CH4+2 02 = C02+2H20 


2.47 
34.8 


Methane 

Nitrogen 


17.4 


Ethylene 


C2H4+3 02 = 2C02+2H20 


14.9 


Oxygen 

Sulphur 




S + 02-S02 


4.32 


Water vapor 











Mean 

Specific 

Heat. 


Density and Specific Volume 

at 32 deg. fahr., and 14.7 Lb. 

per Sq. In.f 


Heat of Combustion (Total 
Heat Value) B.t.u.t 




Weight per 
Cu. Foot. 


Cu. Feet per 
Pound. 


Per Pound. 


Per Cu. Foot 
at 32 deg. fahr. 
and 14.7 Lbs. 


Acetylene 


d 



0.0725 
0.0807 


13.79 
12.39 


21,430 


1582 


Air 




Ash 






Carbon 


145 (solid) 

145 (solid) 

0.1227 

0.0781 

0.0056 

0.0447 

0.0783 

0.0795 

0.0892 

125 (solid) 




4,380 
14,540 




Carbon 






Carbon dioxide. . . 


8.15 
12.80 
177.9 
22.37 
12.77 
12.80 
11.21 




Carbon monoxide . 

Hydrogen 

Methane 

Nitrogen 


4,380 
62,000 
23,840 


342 

345 

1067 


Ethylene 

Oxygen 


21,450 


1685 


Sulphur. . . . 


4.000 




Water vapor 













Atomic. 



t Smithsonian tables. 



t Compiled from various sources 



44 STEAM POWER PLANT ENGINEERING 

For all practical engineering purposes the use of the exact values of 
the molecular weights is an unnecessary refinement and the decimal 
factors may well be omitted. In the ensuing calculations only the ap- 
proximate values will be considered.) If the products of combustion 
are condensed and their temperature lowered to the initial temper- 
ature of the constituent gases the heat liberated will be 62,000 B.t.u. 
This is known as the total heating value. If the products of combustion 
are not condensed, which is the usual case in practice, the latent heat 
of vaporization of the water vapor is not available. The difference 
between the higher heating value and the unavailable heat is called 
the net heating value. The unavailable portion of the heat depends 
upon the temperature at which the products of combustion are dis- 
charged. This varies with practically every installation. Thus, 
one pound of water vapor escaping uncondensed in the products of com- 
bustion at temperature ^i deg. fahr. will carry away approximately 
(1090.6 + 0.46 ti — t) B.t.u. above initial temperature t deg. fahr. of 
the constituent gases. (See paragraph 30.) Since one pound of hydro- 
gen burns to approximately 9 pounds of water vapor, the lower heating 
value h' will be 

h' = 62,000 - 9 (1090.6 + 0.46 ti - t) B.t.u. (4) 

Many attempts have been made to adopt a standard lower heating 
value, but the results have been far from harmonious. The U. S. 
Bureau of Standards recommends "that the quantity to be subtracted 
from the gross value to give the net value be taken as the latent heat of 
vaporization at zero degrees centigrade, of the water formed during 
combustion, and of that contained in the fuel." 

This would give the net or lower heat value of hydrogen as 
62,000 - 9 (1073.4) = 52,340 B.t.u. 
For ti = t = deg. cent. = 32 deg. fahr., formula (4) gives the same 
result. 

Combustion of Bituminous Coals. — H. Kreisinger, Prac. Engr., Apr. 15, 1917, 
p. 347. 

20. Calorific Value of Coal. — The heat liberated by the complete com- 
bustion of unit weight of fuel is called the heating value or calorific value 
of the fuel. The only accurate method of determining this quantity 
for a solid fuel is to burn a weighed sample in an atmosphere of oxygen 
in a suitable calorimeter. An alternative method is to calculate the 
heating value from the ultimate analysis. Approximate results may 
be obtained from empirical formulas based upon the proximate analysis. 

Dulong's formula is the generally accepted rule for calculating the 
heating value of coal. It is based on the assumption that all the oxygen 
in the fuel and enough hydrogen to unite with it is inert in the form of 



FUELS AND COMBUSTION 45 

water and that the remainder of the hydrogen and all of the carbon 
and sulphur are available for oxidation thus: 

hd = 14,600 C + 62,000 (h - ^) + 4000 S * (5) 

in which hd = heating value in B.t.u. per pound of fuel. 

Cj H, and S refer to the proportion by weight of carbon, hydrogen, 
oxygen and sulphur in the fuel. 

Heating values calculated by means of Dulong's formula fail to check 
with calorimetric determinations because 

(1) The heating values of the elements, carbon, hydrogen and sul- 
phur are not accurately estabUshed and the true values may depart 
somewhat from those given in the formula. 

(2) The heating value of an element in the free state is not necessarily 
the same as when a component of a chemical compound, because of ab- 
sorption or evolution of heat during formation of the compound. 

(3) The oxygen content in the ultimate analysis is determined by 
difference. This method throws the sunomation of all the errors in- 
curred in the other determinations upon the oxygen. Furthermore, 
the assumption that all of the oxygen is combined with hydrogen to 
form water is not true since some of the oxygen may be combined with 
carbon. 

However, in spite of these objections, extensive investigations show 
that Dulong's formula gives results which agree substantially with 
calorimetric determinations for all ordinary coals. With lignite, wood 
and other fuels high in oxygen and with some fuels high in hydrogen 
such as cannel coal, the results are not reliable and may be considerably 
in error. 

Numerous attempts have been made to establish empirical formulas 
for calculating the heat value from the proximate analysis but the 
results have been decidedly discordant. Many of these rules give con- 
sistent results when applied to certain classes of fuels or to fuels from 
a given district, but as general laws they may lead to serious error. 

In this connection may be mentioned the investigations of Mahler, f 
Lord and Haas,| Parr and Wheeler,§ Goutal,|| and Kent.1[ 

* In the fuel bulletins of the U. S. Geological Survey and the Bureau of Mines, 
Dulong's formula is stated: 

hd = 14,544 C + 62,028 (h - ^^ + 4050 S. 

t Kent, Steam Boiler Economy, 1915, p. 143. 

J Transactions of American Soc. of Mechanical Engineers, Vol. 27, 1897, p. 259. 

§ Illinois University Engineering Experiment Station, Bulletin 37, 1909. 

II Comptes Rendus de L'Academie des Sciences, Vol. 135, p. 477. 

H Transactions of American Soc. of Mechanical Engineers, Vol. 36, 1914, p. 189. 



46 



STEAM POWER PLANT ENGINEERING 



When a series of tests is being made with a view of improving effi- 
ciency it is of considerable importance to have the results of each test 
immediately after completion of the run in order that the information 
gained may be used in the succeeding tests. For this reason it is par- 
ticularly desirable to determine the heating value of the coal and 
'^ cinders" with as little delay as possible. If the source of the coal sup- 
ply is known the simplest, and a fairly accurate method, is to assume 
a fixed heat value for the combustible. This may be obtained from 
results of previous tests or from results published by the Bureau of Mines. 
For example, the average heat value of the combustible for a number 
of Illinois coals as compiled from Government reports and other sources, 
is 14,300 B.t.u. per pound. With the exception of a very few samples 
the actual heating value varied less than 2 per cent from this average 
and the maximum departure did not exceed 3 per cent. Extensive 
experiments conducted in the power plant laboratory of Armour & 
Company, Chicago, Illinois, show that the heat value of the combustible 
in the refuse or clinkers is practically that of the combustible in the fuel, 
averaging 14,100 B.t.u. per pound for Illinois coals. 

The heating value of any fuel may be determined from the proximate 
analysis with a fair degree of accuracy by calculating the ultimate 
analysis, as shown in the preceding paragraphs, and apnl . inr- ^ulong's 
formula. 

Calorimetric determinations are necessary in all cases where ac- 
curacy is required. 

Example 5. Approximate the heat values for the Illinois coal (analy- 
sis as in Example 1) from the calculated ultimate analysis. 



1. Assuming a fixed heat value for the combustible 

h = 14,300 X 0.8163 

2. Calculated from Dulong's formula: 

* {a) h = 14,600 X 0.65 + 62,000 X 0.0326 + 4000 

X 028 

* (b) h = 14,600 X 0.682 '+62,000 

(o.om-^-^) 

* (c) h = 14.600 X 0.6655 + 62,000 

^0.0428 - ^^^)+ 4000 X 0.0197 . . 

3. Actual value from calorimeter test 



B.t.u. per 
Pound of Coal 
as Received. 


Departure from 

Calorimeter 

Determinations, 

Per Cent. 


11,674 


-2.36 


11,623 


-1.96 


12,053 


+0.80 


11,869 


-0.76 


11,957 


0.00 



(a) Ultimate analysis calculated from average analysis of Illinois coals. See Example 4, 

(b) Ultimate analysis calculated from proximate analysis (Equations (1) to (3) ). 

(c) Ultimate analysis from chemical tests. 



FUELS AND COMBUSTION 



47 



TABLE H. 

VARIATION IN CALORIFIC VALUE OF FUELS. 

(As Mined.) 



Air-dried wood 

Air-dried peat 

Lignite 

Sub-bituminous coal 

Bituminous coal 

Semi-bituminous coal 

Anthracite coal 

California crude oil 

Pennsylvania heavy crude oil 



B.t.u. 



6,000 to 


7,500 


About 


7,500 


5,200 to 


7,500 


5,500 to 11,500 


10,000 to 14,500 


13,500 to 14,900 


11,000 to 13,800 


17,000 to 19,300 


About 


20,700 



21. Air Theoretically Required for Complete Combustion. — The com- 
bustible portion of all commercial fuels consists chiefly of carbon and 
hydrogen and a small percentage of volatile sulphur. Based upon the 
approximate molecular weights the carbon, hydrogen and sulphur 
require the following weights of oxygen for complete combustion: 



1 lb. carbon requires 



-^^ = j2 = 2.66 + lb. oxygen. 



O 32 
1 lb. hydrogen requires pr^ = — = 8.00 lb. oxygen. 

I xl2 4 

O 32 
1 lb. sulphur requires "o" ^ oo "= l-^^ ^t>. oxygen. 



In the ordinary furnace the oxygen is obtained from the atmosphere 
which, neglecting moisture and a few minor elements, contains oxygen 
and nitrogen mechanically mixed as follows: 

PROPORTION OF NITROGEN AND OXYGEN IN DRY ATMOSPHERIC AIR. 



Nitrogen. . . . 

Oxygen 

Nh-0 

(N -f- 0) 4- O 



Exact Value. 



By Volume. By Weight 



79.09 
20.91 
3.782 

4.782 



76.85 

23.15 

3.32 

4.32 



Approximate Value. 



By Volume. 



79.0 
21.0 
3.76 
4.76 



By Weight. 



77.0 
23.0 
3.34 
4.35 



48 STEAM POWER PLANT ENGINEERING 

Hence the dry air requirements are: 

1 lb. of carbon requires 2.66 X 4.35 = 11.58 lb. dry air. 
1 lb. of hydrogen requires 8.00 X 4.35 = 34.8 lb. dry air. 
1 lb. of sulphur requires 1.00 X 4.35 = 4.351b. dry air and for a com- 
pound fuel 

Ai = 11.58 C + 34.8 ("h - §) + 4.35 S, (5) 

in which 

Ai = weight of dry air required. 

C, H, 0, and S = proportional part of the carbon, hydrogen, oxygen 
and volatile sulphur in the fuel. 

-5- = proportional part of the hydrogen supplied with oxygen from the 

fuel itself.* 

It should be borne in mind that these values are based on the ap- 
proximate molecular weights of the various elements and the assumption 
that the air is composed of 23 parts oxygen and 77 parts of nitrogen, 
by weight. Using the exact molecular weights, as fixed by the Inter- 
national Committee on Atomic Weights, and taking the air as composed 
of 23.15 per cent oxygen and 76.85 per cent nitrogen, equation (5) be- 
comes 

Ai = 11.5 C + 34.2 ^H - ^) + 4.3 S. (6) 

In using equation (6) in connection with the determxination of heat 
losses, to be consistent, all calculations should be made with the exact 
molecular weights and the true ratio of nitrogen to oxygen in atmospheric 
air. The theoretical weights of air as calculated from equations (5) 
and (6) differ by approximately one per cent as a maximum. 

Example 6. Required the theoretical weight of dry air supplied per 
pound of coal as fired with analysis as follows: 

Per Cent. Per Cent. 

Carbon 65 Ash and Sulphur 13 

Hydrogen 5 Water 8 

Oxygen 8 Total 100 

Nitrogen 1 

* This term ^ H — -^ j does not contain a proper correction for the hydrogen con- 
tained in the moisture, for not all of the oxygen in coal is combined with hydrogen. 
Part of the oxygen is probably combined with nitrogen in organic nitrates and par^ 
may be present in carbonates in mineral matter caught in the coal. The error dS 
this assumption, however, lies within the accuracy of the average boiler observations. 



FUELS AND COMBUSTION 49 

Substituting the value of C, H, and in equation (5) 

Ai = 11.58 X 0.65 + 34.8 ^0.05 - ^\ = 8.92 pounds, 

the theoretical weight of dry air necessary to burn one pound of coal 
as fired. 

Since the coal contains 8 per cent of moisture the weight of dry air 
required per pound of dry coal is 

jr-^ = 9.69 pounds. 

The water and ash only are treated as incombustible, therefore the 
air required per pound of combustible is 

^-^ = 11.29 pounds. 

Similar calculations for different fuels will show that the theoretical 
air requirements per pound of fuel or combustible vary within wide 
limits. When expressed in terms of theoretical air requirements per 
10,000 B.t.u., however, there is a close agreement between all fuels. 
Several hundred fuels ranging from peat to crude oil rated on this basis 
gave an average value of 7.5 pounds of air per 10,000 B.t.u. with a 
maximum departure not exceeding 2 per cent. The calorific value of 
the coal in the preceding example is 15,150 B.t.u. per pound of com- 
bustible; on the B.t.u. basis this gives 

j^^X7.5 = 11.35 pounds, 

which checks substantially with the calculated value. See also Table 13. 
22. Products of Combustion. — A knowledge of the constituents of the 
solid and gaseous products resulting from the combustion of a fuel offers 
a means of determining the losses incident to such combustion. For 
maximum efficiency complete combustion with theoretical air re- 
quirements is necessary and the resulting products should consist only 
of CO2, N2, H2O, ash, and the oxides of other combustible elements in 
the fuel. The dry gaseous products, such as appear in the commercial 
flue gas analysis, will consist of CO2 and N2 only, since the SO2, if there 
is any, is partly absorbed by the water in the sampling apparatus 
(see paragraph 415) while some of it probably goes into the CO2 pipette 
and appears in the analysis as CO2. It is difficult to determine the 
exact distribution but since the maximum error due to this source does 
not exceed 0.2 per cent it is common practice to disregard the SO2 
entirely. If combustion is complete but air is used in excess of theo- 
retical requirements the gaseous products will include free oxygen. If 
combustion is incomplete CO will also be present in the gaseous products 
and perhaps small quantities of hydrocarbons. The following ex- 



50 



STEAM POWER PLANT ENGINEERING 



amples illustrate some of the accepted methods for determining the 
constituents of the products of combustion. 

Example 7. Required the character and amount of the products of 
combustion if one pound of coal, as per following ultimate analysis, 
is completely burned with theoretical air requirements. 



Carbon 65 Ash . . . . 

Hydrogen 5 Water . . 

Oxygen 8 Sulphur . 

Nitrogen 1 Total . . . 



12 

8 

1 

100 



TABULATED CALCULATIONS. 





Pounds of Substance per Pound of Coal as Fired. 




Co 


H2 


O2 


N2 


CO, 


HoO 


Ash. 


The carbon will produce: 

Carbon 


0.65 














65 X ft 








2.38 






0.65 X f i 






1.73 








. 65 X ft X H 






5.80 








The available hydrogen will 
produce: 
Hydrogen 




0.04 










(0.05 "f^g 










0.36 




C0.05 "f^8 






0.32 








V 8 )" 

Co.05 ''"^'Isxli 






1.07 








The oxygen and inert hydrogen 
will produce: 
Hvdroeen 




0.01 










Oxygen 




0.08 










0.08 + Of 










0.09 




' 8 
The nitrogen in the fuel* is con- 
sidered inert 








0.01 






The moisture will appear as 
vapor 










0.08 




Ash plus sulphur f 












0.13 


Total 


0.65 


0.05 


2.13 


6.88 


2.38 


0.53 


0.13 











* This is not strictly true since a portion of the nitrogen content of the fuel appears in the flue gas in 
combination with other elements, but the amount is so small compared with that supplied in the air 
that no appreciable error arises from the assumption that it remains inert and passes through the furnace 
without change. 

t The sulphur content is ordinarily so small that no attempt is made to separate the volatile and non- 
volatile constituents and the whole is treated as ash. If the volatile portion is to be considered the in 
fluence of the SO2 or SO3 in the flue gas should be included in the heat balance. Some engineers treat 
one-half the sulphur as volatile and the balance as ash. 

Total gaseous products = CO2 + N2 + total H2O 

= 2.38 + 6.88 + 0.53 = 9.79 lb. 

Or, separating the compounds into their elementary constituents. 
Total gaseous products = C + H2 + O2 + N2 + free H2O 
= 0.65 + 0.05 + 2.13 + 6.88 + 0.08 
= 9.791b. 



FUELS AND COMBUSTION 51 

Total dry gaseous products = 9.79 - 0.53 (total H2O) = 9.26 lb. 
Dry air = 9.26 - C + 8 ^H - ^ W N2 (in the fuel) 

= 9.26 - 0.65 + 8^0.05 - ^] - 0.01 = 8.92 lb., 

which checks with the results as calculated from equation (5). Since 
the dry air consists of all the nitrogen supphed by the air and the oxygen 
for the combustion of the carbon and hydrogen we have as an addi- 
tional check, 

Dry air = 6.87 + 0.65 X j? + 8 (o.05 - ^) 
= 6.87 + 1.73 + 0.32 = 8.921b. 

If the coal in the preceding problem is completely burned with 33J 
per cent air excess the products of combustion will be the same as before 
with the exception of the addition of J X 8.92 = 2.97 lb. dry air. The 
gases, by weight, will consist of CO2 = 2.38 lb., N2 = 1| X 6.87 + 0.01 
= 9.17 lb., free O2 = 4 X 2.05 = 0.68 lb., H2 = 0.53 lb. or a total of 
12.76 lb. Weight of dry gases per lb. of coal = 12.76 - 0.53 = 12.23 
lb. 

The free oxygen comes from the air supplied and not used. This 

79.01 
oxygen is accompanied by ' = 3.78 times its volume of nitrogen. 

(N — 3.78 0) represents the nitrogen content of the air actually re- 
quired for the combustion represented by the flue gas analysis. Hence 

N 
N _ Q 7Q o ^^ ^^® ^^^^^ ^^ ^^^ ^^^ supplied to that theoretically required 

to burn the coal to CO and CO2. For the example under consideration 

N 81.2 



N- 3.78 81.2-3.78 X 5.4 



= 1.335. 



100 X 1.335 — 100 = 33.5 per cent = air excess, which agrees substan- 
tially with results as previously determined. * ' 

If all the carbon had burned to CO2 the ratio of the total air supply 
to that theoretically required for complete combustion is 

N 

N - 3.78 (0 - i CO) ' 

N in this case represents the nitrogen incident to the complete com- 
bustion of the carbon; (O — ^ CO) represents the equivalent volume ot 
oxygen due to air excess since carbon combines with one volume of 
oxygen to form two volumes of CO. 

It will be noted that dry air only has been considered in the foregoing 
calculations. Atmospheric air is never dry, hence the weight of volume 
of atmospheric air will differ from the amounts as calculated above. 
For most engineering purposes atmospheric air may be considered dry. 



52 STEAM POWER PLANT ENGINEERING 

For methods of determining the weight of dry air in atmospheric air 
see paragraph 470. 

In the preceding calculation the products of combustion have been 
expressed on a weight basis, which, as will be shown later, is most con- 
venient for calculating the various heat losses. However, in deter- 
mining the gaseous constituents of the products of combustion the 
measurements are made volumetrically. The transfer from one basis 
to the other is readily effected by the following adaptation of Avo- 
gadro's law : * 

Lb. per cu. ft. of any gas = 0.00278 m. (7) 

Conversely, cu. ft. per lb. of any gas = 358.6 -^ m, (8) 

in which 

m = molecular weight of the gas referred to oxygen as 32. Volumes 
measured at 32 deg. fahr. and atmospheric pressure, 29.92 
inches of mercury. 

Thus, the volume of one pound of CO2 at 32 deg. fahr. and 29.92 
inches of mercury = V = 358.6 -J- 44 = 8.15 cu. ft. 

Similarly the volumes of one pound of oxygen and nitrogen are 11.21 
and 12.77 cu. ft. respectively. 

Example 8. Transfer the flue gases in Example 7 from a weight to a 
volume basis. 

In Example 7 it was shown that for complete combustion with theo- 
retical air requirements the dry gaseous products of combustion con- 
sisted of 2.38 lb. CO2 and 6.88 lb. N2. 

2 38 — 44 
Percent CO. by vol. = 100 X 238^44 + 6.88^28 = ^^'^^ 

For complete combustion with 33^ per cent air excess the dry gaseous 
products consisted of 2.38 lb. CO2, 9.16 lb. N2, and 0.69 lb. free O2. 
Transferring to the volumetric basis: 

2 38 — 44 
Per cent CO2 by vol. = 2.38 -^ 44 + 9". 16 -I 28 + 0.69 -^ 32 "" ^^^ 

^ 100 X Q-Q54 ^ 5.4 ^ 

0.054 + 0.327 + 0.022 0.403 

32 7 
Per cent N2 by vol. = ^-^ = 81. 1. 

2 2 
Per cent free O2 by vol. = ^ '^ . = 5.5. 

U.Uo4 

* Equal volumes of all gases contain the same number of molecules when at the 
same temperature and pressure. 



FUELS AND COMBUSTION 



53 



When chemical reactions are expressed in terms of molecules the 
coefficients of the molecule symbols represent relative volumes, thus, 
the reaction C + O2 = CO2, shows that one volume of oxygen combined 
with carbon forms one volume of CO2, both being measured at the same 
temperature and pressure. Therefore, the volume of CO2 after com- 
bustion is precisely the same as that of the oxygen before it was 
combined with the carbon. The volume of one pound of CO2 as de- 
termined above is 8.15 cubic feet and since one pound of carbon unites 
with 2§ pounds of oxygen to form 3f pounds of CO2 we have 



Substance. 



Weight per Lb. of 
Carbon, Lb. 



Spec. Volume, 
Cu. Ft. per Lb. 



Actual Volume 
Resulting from 
the Combustion 
of 1 Lb. of Car- 
bon, Cu. Ft. 



Per Cent by 
Volume. 



For Theoretical Air Requirement. 



(C02 

Free ] . 


31 X 


8.85 X 


8.15 = 


29.89 


20.91 


/n 


12.77 = 


113.01 


79.09 


Total 


142.90 


100.00 











For 50 Per Cent Air Excess. 



(C02 


31 X 
0.5 X2f X 
1.5X8.85 X 


8.15 = 
11.21 = 
12.77 - 


29.89 

14.94 

169.56 


79 09 


Free ^0 

( N. 






Total 


214.39 


100.00 











For 100 Per Cent Air Excess. 



(C02 

Free^O 

(N 


31 X 

2| X 

2X8.85 X 


8.15 = 
11.21 = 

12.77 = 


29.89 

29.89 

226.08 


10.45+;2o 91 
10. 45+1'^^^^ 
79.09 


Total 


285.86 


100 00 











It will be noted that the actual volume of CO2 is always the same 
irrespective of the excess of air supplied, while the percentage by volume 
decreases as the excess of air increases. In each case CO2 + = 20.9 
is constant. (The approximate value 21 is ordinarily taken instead of 
the exact quantity, 20.9.) 



54 



STEAM POWER PLANT ENGINEERING 



The actual volume of oxygen and the percentage by volume increase 
with the amount of excess air, therefore either the CO2 or O content of 
the products of combustion is a true index of the air excess. This 
apphes to the complete combustion of pure carbon only. Assuming an 
average theoretical air requirement of 11.5 pounds of air for the complete 
combustion of pure carbon the resulting air requirements for different 
percentage of CO2 are given in Table 12. Although the actual volume 
of nitrogen increases with the air excess its volume percentage remains 
the same after combustion as before. The nitrogen performs no useful 
function in combustion and passes through the furnace without change. 
It simply dilutes the oxygen for combustion and its presence in the 
flue gases represents a large percentage of the heat lost in the chimney. 

CO produced by incomplete combustion of carbon will occupy twice 
the volume of oxygen entering into its composition as is evidenced from 
the molecular reaction 2C + O2 = 2C0. 

Therefore, with pure carbon as fuel, the sum of the percentages by 
volume of CO2, O2 and J CO must be in the same ratio to the nitrogen in 
the flue gas as is oxygen to the nitrogen in the air supplied; viz., 20.91 
to 79.09. When burning coal, however, the percentage of nitrogen is 
obtained by subtracting the sum of the percentages by volume of the 
other gases from 100. 

In commercial furnace practice CO2 is used as the index to efficiency 
of combustion because of the ease with which it is obtained. For 
fuels high in volatile matter the per cent of CO2 in the products of com- 
bustion is less than 20.91 for complete combustion, since the oxygen 
which combines with hydrogen to form H2O does not appear in the 
sample as ordinarily tested : thus for heavy crude oil the corresponding 
maximum content of CO2 is approximately 16 per cent. The air re- 
quirements and resulting CO2 content for complete combustion of a 
number of typical fuels are given in Table 13. 



TABLE 12. 

WEIGHT OF AIR PER POUND OF CARBON AS INDICATED BY THE PERCENTAGE OF 

CO, IN THE FLUE GAS. 



Per Cent of 
CO,. 


Pounds of Air. 


Per Cent of 
CO2. 


Pounds of Air. 


Per Cent of 
CO,. 


Pounds of Air. 


20.9 


11.5 


14 


17.1 


7 


34.3 


20 


12.0 


13 


18.5 


6 


40.0 


19 


12.6 


12 


20.0 


5 


48.0 


18 


13.3 


11 


21.8 


4 


60.0 


17 


14.1 


10 


24.0 


3 


80.0 


16 


15.0 


9 


26.7 


2 


120.0 


15 


16.0 


8 


30.0 


1 


240 



FUELS AND COMBUSTION 



55 



TABLE 13. 

THEORETICAL AIR REQUIREMENTS FOR VARIOUS FUELS AND THE RESULTING 
MAXIMUM PER CENT CO2 IN THE FLUE GAS FOR COMPLETE COMBUSTION. 



Fuel, Moisture and 
Ash Free. 



Pure carbon 

Anthracite 

Semi-anthracite. . 
Semi-bituminous. 

Bituminous 

Sub-bituminous. . 

Lignite 

Peat 

Crude oil 





Ultimate Analysis.* 




Air, Pounds. 


C. 


H. 


N. 


0. 


s. 


Per Pound 
of Fuel. 


Per 10,000 
B.t.u.t 


100.00 
94.39 










11.58 

11.39 




1.77 


0.71 


2.13 


1.00 


7.4 


89.64 


3.97 


0.63 


3.23 


2.53 


11.59 


7.5 


86.39 


4.84 


1.46 


5.50 


1.81 


11.41 


7.6 


79.71 


5.52 


1.52 


9.87 


3.38 


10.70 


7.4 


78.06 


5.70 


1.35 


13.10 


1.79 


10.24 


7.3 


70.64 


4.61 


1.22 


22.67 


0.86 


8.75 


7.3 


59.42 


5.50 


1.50 


33.33 


0.25 


7.30 


7.6 


84.90 


13.7 


0.60 


0.80 




14.45 


7.5 



C02, 

Per Cent 
by Volume. 



20.91 
20.06 
20.00 
18.65 
18.46 
18.56 
19.68 
20.79 
15.90 



* Compiled from Bulletins No. 22 and No. 85, U. S. Bureau of Mines. 

t 200 samples of various fuels gave an average theoretical air requirement of 7.5 pounds per 10,000 
B.t.u. (Bomb calorimeter). The maximum variation did not depart more than 2 per cent from the 
average value. 

In coal-burning practice, from 15 to 16 per cent of CO2 is all that can 
be expected under the very best conditions, with an average range 
for general practice between 10 per cent and 14 per cent. Anything 
less than 12 per cent shows an excessive amount of air supplied. Trav- 
eUng grates, unless carefully operated, are apt to show as low as 5 
per cent of CO2. 



15 

a 13 



10 



































- 




















.,, 
























































— 















































" 




































c^S 


— - 




-^ 




































, 




— 1 


^ 


— ' 






































, 





— 


"^ 














Relation of Gas Composition in Rear 

Combustion Chamber To Temperature 

at Same Place 


- 






















• 






-^ 






























































■~^ 


^ 






























































^■ 


-^ 


^ 








Sj 




















































^ 




























y 




































■~~- 




■^ 




^ 










y 




















































^ 




— 


— 












































^ 


































__ 




_C 


=d 


— 




— 


-~" 






















L 




















~ 








~ 






" 























d900 sooo 



s 

s 
0.5 o 

0.35. 

0.20 

0.1** 



2700 



0. 3 

2800 8 



Fig, 



2100 2200 2300 M)0 2500 2600 
CombusUoji Gliamber Temperatiire.Deg.Fah. 

8. Relation of Gas Composition in Combustion Chamber to Temperature 



It must not be assumed that a high percentage of CO2 in the flue 
gas is necessarily a true indication of good combustion and hence of 
high efficiency. As the percentage of CO2 increases there is a tendency 
for the CO to increase also (see Fig. 8) and the thermal gain due to 



56 STEAM POWER PLANT ENGINEERING 

minimum air excess as indicated by the high percentage of CO2 may be 
more than offset by the loss due to incomplete combustion (see para- 
graph 27). Determinations of the CO content of the flue gas are neces- 
sary for an accurate heat balance and particularly so if the CO2 content 
is high. 

33. Air Actually Supplied for Combustion. — In practice the amount 
of air supplied is measured directly in situations where such measure- 
ments can be readily made, as in connection with mechanical draft, or 
where the entire air supply is forced to flow through a conduit. In 
most cases, however, physical measurements of flow are not feasible 
and the amount of air supplied is calculated from the flue-gas analysis.* 
The latter offers a fairly accurate method for determining air excess, 
provided the sample of gas is truly representative of average conditions. 

Based upon the ratio of the combining weights, the weight of carbon 
in CO2 = t\ CO2, and that in CO = f CO. If CO2 + O + CO + N = 
total gas in percentage by weight the weight A3 of the dry gas per pound 
of carbon actually burned is 

^ CO2 + O + CO + N 
' t\C02 + f CO * ^ ^ 

Multiplying each gas by its respective molecular weight, viz., CO2 = 
44, O = 32, CO and N = 28, and reducing, we have 

^ llC02 + 80 + 7(CO + N) 
' 3(C02 + CO) ' ^^^^ 

in which 

A3 = weight of dry gas per pound of carbon actually burned. 

CO2, CO, O and N = percentages by volume of the carbon dioxide, 
carbon monoxide, oxygen and nitrogen in the flue gas. 

Since CO2 + CO + O + N= 100, neglecting traces of minor con- 
stituents, CO = 100 - CO2 - O- N. Substituting this value of CO 
in equation (10) and reducing, we have 

4 CO2 + O + 700 



3 (CO2 + CO) 



(11) 



Example 9. Determine the weight of dry air supplied per pound of 
coal as fired, analysis as in paragraph 22, if the flue gas resulting from 
the combustion is composed of 

CO2 12.8 per cent by volume. 

CO 0.6 per cent by volume. 

O2 5.4 per cent by volume. 

N2 81.2 per cent by volume (by difference). 



* For Flue-Gas see Par. 415. 



FUELS AND COMBUSTION 67 

Substituting the various percentages in equation (11) 

A3 = o MO g I (\r\ '"^ 18.82 lb. ot dry gas per lb. of carbon 

actually burned. 

Since the coal as fired contains 0.65 carbon, the dry gas per lb. of 
coal burned = 18.82 X 0.65 = 12.23 lb. If part of the coal falls 
through the grate, as is always the case in practice, the weight of carbon 
actually burned should be taken instead of the total carbon content. 

The total weight of dry air actually supplied per pound of coal burned 

IS / 09\ 

12.23 - 0.65 + 8(0.05 - ^j = 11.90. 

It has been previously shown (paragraph 21) that the coal under 
consideration requires 8.92 pounds of air for theoretical combustion, 

^'""'^ A- ,nn 11-90 - 8.92 „„ . 

Air excess = 100 = 33.4 per cent. 

o.vZ 

The 7 N in equation (10) represents the N supplied by the air less the 
negligible amount furnished by the fuel itself. Since the nitrogen con- 
tent of air is 77 per cent of the weight of the air, we have 

= 7N _ A 77 = 3.03 N 

' 3 (CO2 + CO) • CO, + CO' ^^^^ 

in which 

A4 = the weight of dry air supplied per pound of carbon burned. 
N, CO2, CO = percentages by volume of nitrogen, carbon dioxide and 
carbon monoxide in the flue gas. 

For the example cited above 

. 3.0 3 X 81 .2 ^^_ , 

^'= 12.8 + 0. 6 ^^^-^^P^^"^^- 

For the coal under consideration 

Dry air per pound = 0.65 X 18.36 = 11.93. 

This checks practically with results calculated from equation (11). 

The relation between excess air and CO2 in the flue gases for a specific 
case is illustrated in Fig. 9. These results were obtained from a 508 
horsepower Babcock & Wilcox boiler equipped with chain grate and 
burning Ilhnois coal. (University of Ilhnois Bui. 32, April 12, 1915.) 

Air Excess in Boiler Furnace Practice: National Engr., Feb., 1915, p. 90. 

The Importance of CO^, as an Index to Combustion and in Connection with Flue 
Gas Temperature, to Boiler Efficiency: Trans. A.S.M.E., 32-1215. Flue Gas Analysis 
and Calculations: Power, Aug. 9, 1910; Eng. Review, Aug., 1910. Real Relation 
of CO2 to Chimney Losses: Power, Dec. 7, 1909, p. 969. Sampling and Analysis of 
Furnace Gas: Power, Aug. 22, 1911, p. 282; Bulletin No. 97, U. S. Bureau of Mines, 



58 



STEAM POWER PLANT ENGINEERING 










s. 


\^ 
















1 




















\ 


4.^^ 


\, 










O From Analysis at Furnace 
« From Analysis at Flue 
















19\ 


Ks 


k 










Actual Curve 
Theoretical Curve 
























'\\ 


^ 


$^J 




































3 


o^> 


^' 


\ 


^ 






































2> 


^ 


5^ 




































a 




^^ 


< 


^ 


»i! 








































— 


T^ 


9 


.^_^ 








































i 

















































10 20 30 40 50 60 70 80 90 J 00 110 120 130 140 150 160 170 180 190 200 
Excess Air (Percent) 

Fig. 9. Relation between Excess Air and CO2 in Flue Gases. 

24. Temperature of Combustion. — The actual temperature incident 
to the combustion of a fuel is most satisfactorily determined by means 
of a suitable thermometer or pyrometer. The theoretical temperature of 
combustion may be calculated from the simple relationship 



ti= — + t, 
ws 



(13) 



in which 



ti = final temperature of the products of combustion, deg. fahr. 
h = low calorific value of the fuel, B.t.u. per pound. 
s = mean specific heat of the products of combustion. 
w = weight of the products of combustion, pounds per pound of fuel. 
t = initial temperature of the fuel and air supply, deg. fahr. 

Thus, in the combustion of one pound of carbon with theoretical air 
requirements, initial temperature 62 deg. fahr., the maximum theoretical 
temperature will be 



^1 



14,540 



12.58 X 0.29 



+ 62 = 4000 deg. fahr. (approx.). 



No such temperature has ever been obtained in practice from the 
combustion of carbon in air. The discrepancy between actual and 
calculated results is attributed to (1) difficulty of effecting complete 
combustion with theoretical air supply, (2) radiation losses, (3) error in 
the assumed value of the mean specific heat at this high temperature, 
and (4) uncertainty of the proportion of the calorific value of the fuel 
available, at this high temperature, for increasing the temperature of 
the products of combustion. 



FUELS AND COMBUSTION 59 

An inspection of equation (13) will show that the greater the weight 
of the products of combustion for a given weight of fuel, the lower will 
be the temperature of combustion. Evidently, for maximum tempera- 
ture the weight of air supplied per pound of fuel should be kept as low 
as possible, consistent with complete combustion. A perfect union 
of fuel and air in theoretical proportions is almost impossible, and to 
insure complete combustion an excess of air is necessary. The influence 
of air dilution on temperature of combustion is best illustrated by a 
practical example: 

Example 10. Required the theoretical temperature of combustion 
of carbon in air if 50 per cent air excess is necessary for complete com- 
bustion. Since the complete oxidation of one pound of carbon requires 
11.58 pounds of air, the weight of the products of combustion will be 
11.58-1-0.5 X 11.58-1- 1 = 18.37 pounds and the final increase in 
temperature will be 

14 540 
ti = 18 37 X 27 ^ ^^^^ ^^^' ^^^^' (^PP^^^-)- 

Data relative to the specific heats of gases are rather discordant. 
The following equations are considered by the U. S. Bureau of Standards 
to be as nearly accurate as it is possible to give at the present time (1916). 

For N2 s = 0.249 + 0.000^019 t ■ (14) 

CO s = 0.250 4- 0.000'019 t (15) 

O2 s = 0.218 -^ 0.000^017 t (16) 

H2 s = 3.40 + 0.000'27 t (17) 

Air s = 0.241 -f 0.000'019 t (18) 

CO2 5 = 0.210 + 0.000^0742 t - 0.000'000'018 t^ (19) 

in which 

s = mean specific heat at constant atmospheric pressure and tem- 
perature range deg. cent, to t. 
t = maximum temperature. 

For the mean specific heat of H2O vapor see paragraph 449. 

Between 1000 deg. cent, and 1500 deg. cent, the results are un- 
certain and dependence can be placed in only the first two significant 
figures in the decimal. Beyond 1500 deg. cent, the results are purely 
conjectural since experiments have not been made at these high tem- 
peratures. The values of the mean specific heat (s = 0.29 and s = 0.27) 
used in the preceding computations were calculated from these equations. 
The value s = 0.27 is probably not far from the truth, but the value 
s =0.29 may be considerably in error. 



60 



STEAM POWER PLANT ENGINEERING 



The mean specific heat between any two temperatures ti and t may be 
determined by substituting (^i + t) for t in above equations. 

If the mean specific heats, Si, S2 . . . Sn, and weights, Wi, W2 . . . Wn, 
of the constituent gases of a compound are known the mean specific 
heat, s, of the compound may be determined as follows: 



s = 



WiSi + W2S2 



+ WnSn 



Wi + W2 



+ w, 



(20) 



The application of formulas (14)- (19) at high temperatures to equa- 
tion (13) necessitates laborious calculations, and since the results are 
only approximate at the best, extreme refinement in calculation is 
without purpose. The curves in Fig. 10 are plotted from these equations 
and afford a means of approximating the mean specific heat without 
the labor of solving the equations. 









































- 
































1 


-y 


0.28 






































































=* 


=s: 




























































_^;i--=is===^ 










_ 


— 


- 


< 0.27 


















































^ 




" 




- 


, 




1— 


-- 




























/C 


)0 




^ 


-s= 


ss==f^ 


' 










— 




"^ 












8 0.26 




















1^ 


<= 


i== 


''^^.. ' ! 


as. 




^;:<L_ 


---rf^ 




































ss 


=ss 


iss^^^f^^""^ i 






_L— ^=5^^^ 






































<=^0.25 




- 




tss 


tss 










_J — \ — ■ 


" 




^ 


.^ 


■' 










































~ 












J 


^ 


-« 


^ 


r 






! 




^ 












































■^ 





^0.21 

-§0.23 


_- 




j — ' 


















^ 


^ 


























, 


-— 






" 














- 


^ 
























y 














- 




^ 



























- 




-^ 




-— 
















,C0 


^ 


" 








1^ 


.__ 


— 
















_^ 






— *■ 
































^ 


y 


c 






— 


— ' 












H 




" 






— ■ 
































„0.22 






- 


__ 




— ■ 


^ 






















.^ 


-— 








































— 








/ 














L- 


— ■ 


"- 




















































k'-'' 






1/ 




- 


^ 


u 




■ — ■ 


" 




























































/ 


-r' 


^ 


— - 








































































§ 
















































































®0.20 


















































































0.19 
































































































































































































. 





















































100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2C0O 

Temperature, Degrees Centigrade 
32 212 392 572 752 932 1112 1292 1472 1652 1832 2012 2192 2372 3552 2732 2912 3092 3272 3452 3632 

Temperature, Degrees Fahrenheit 

Fig. 10. Mean Specific Heats of Gases at Constant Pressure. 



25. Heat Losses in Burning Coal. — A boiler in order to entirely 
utiUze the heat of combustion of the fuel must be free from radiation 
and leakage losses, the fuel must be completely oxidized and the prod- 
ucts of combustion must be discharged at atmospheric temperature. 
Commercially such conditions are unobtainable, hence complete utili- 
zation of the heat generated is impossible. A boiler which utihzes 
83 per cent of the heat value of the fuel is exceptional and an average 
figure for very good practice is not far from 77 per cent. The various 



FUELS AND COMBUSTION '6l 

losses including the heat utilized by the boiler constitute the commercial 
''heat balance." The losses considered are: 

1. Loss in the dry chimney gases. 

2. Loss due to incomplete combustion. 

3. Loss of fuel through the grate. 

4. Superheating the hygroscopic moisture in the air. 

5. Moisture in the fuel. 

6. Loss due to the presence of hydrogen in the fuel. 

7. Unburned fuel carried beyond the combustion chamber in the 
form of soot or smoke. 

8. Radiation and minor losses. 

Some of these losses are preventable. Others are inherent and can- 
not be avoided. 

26. Loss in the Dry Chimney Gases. — This loss depends upon the 
type and proportion of the boiler and setting and upon the rate of 
driving. It is usually the greatest of all the losses. The heat carried 
away may be expressed: 

hi=Witc-t)c, (21) 

in which 

hi = B.t.u. lost per pound of fuel. ! 

W = weight of dry chimney gases per pound of fuel. (See equation 10.) 
tc = temperature of the escaping gases, deg. fahr. 
t = temperature of the air entering the furnace. 
c = mean specific heat of the dry gases. (This may be taken as 
0.24 for most purposes.) 

It will be noted that the magnitude of this loss depends chiefly upon 
the air dilution and the temperature at which the gases are discharged. 
Flue temperatures less than 450 deg. fahr. are seldom experienced 
except in connection with economizers, and the air dilution is ordinarily 
in excess of 50 per cent of theoretical requirements, hence the loss 
from this cause may range from 8 per cent to 40 per cent of the total 
heat generated. In excellent practice it is not far from 12 per cent with 
a general average of from 20 to 25 per cent. In exceptional cases a loss 
from this cause as low as 9 per cent has been recorded. (Jour. A.S.M.E., 
Nov., 1911, p. 1463.) ~ 

Table 14 indicates the magnitude of the losses for different chimney 
temperatures and weights of air per pound of carbon. 



62 



STEAM POWER PLANT ENGINEERING 



TABLE 14. 
HEAT CARRIED AWAY BY THE DRY CHIMNEY GASES PER POUND OF CARBON 



a 
1 

6 
•s 

3 
3 




Temperature of Chimney Gases. Deg. Fahr. 


300° 


350*> 


400° 


450° 


500° 


660° 


600° 


650° 


12 

* 


750 

5.2 


905 

6.2 


1060 

7.3 


1216 

8.7 


1370 

9.5 


1528 

10.5 


1683 

11.6 


1840 

12.7 


15 


865 

6 


1112 

7.6 


1305 

9.1 


1498 

10.3 


1679 

11.6 


1880 

13.0 


2072 

14.3 


2262 

15.6 


18 


1004 

7.2 


1321 

9.1 


1550 

10.7 


1778 
12.2 


2010 

13.9 


2235 

15.4 


2460 

17 


2692 

17.9 


21 


1266 

8.7 


1530 

10.5 


1785 

12.3 


2060 

14.2 


2320 

16 


2582 
17.8 


2846 

19.5 


3118 

21 


24 

27 


1440 

9.9 


1740 

12 


2040 

14 


2340 

16.1 


2640 

18.2 


2940 

20.3 


3240 

22.4 


3540 

24.4 


1611 

11.1 


1950 

13.5 


2281 

15.7 


2620 

18.1 


2958 

20.4 


3291 

22.7 


3628 

25 


3962 

27.4 


30 


1785 

12.4 


2160 

14.9 


2530 

17.4 


2900 

20 


3270 

22.6 


3641 

25 


4016 

27.8 


4396 

30.4 


33 
36 
39 


1957 

13.5 


2362 

16.3 


2779 

19.2 


3180 

22 


3589 

24.7 


4000 

27.6 


4405 

30.5 


4820 

33.2 

5290 

36.6 


2130 

14.7 


2579 

17.8 


3020 

20.8 


3461 

23.9 


3910 

27 


4350 

30 


4798 

33 


2300 

15.9 


2781 

19.2 


3261 

22.5 


3743 

25.8 


4220 

29.2 


4700 

32.4 


5180 

35.7 


5670 

39 


42 


2479 

17.1 


2999 

20.6 


3508 

24.7 


4023 

27.7 


4540 

31.3 


5052 

34.8 


5570 

39.4 


6100 

42 



* Assumed theoretical requirement. 

Large type gives the loss in B.t.u. per pound of carbon. 

Small type gives the per cent loss, assuming a calorific value of 14,540 B.t.u. pel- 
pound of carbon. 

27. Loss Due to Incomplete Combustion. — If the volatile gases are not 
completely oxidized, as when the air supply is insufficient or the mix- 
ture of air and gases is not thorough, some of the carbon may escape 
as CO. Some of the hydrocarbons may also pass through the furnace 
without being burned. (See Table 15.) The presence of even a small 
amount of CO in the flue gas is indicative of a very appreciable loss, 



FUELS AND COMBUSTION 



63 



TABLE 15. 
ANALYSIS OF CHIMNEY GASES. 
(Report of Committee for Testbg Smoke-preventmg Appliances, Manchester, England, 1905.) 



Boiler. 


Smoky. 


Clear. 


CO2 


O2 


CO 


CH4 


H2 


N2 


COj 


O2 


CO 


CH4 


H3 


N, 


No. 1, hand fired 


(11.00 
ll0.65 


6.90 
6.45 


0.90 
2.15 






81.20 
80.75 






























No. 1, with smoke-pre- 






17.00 
19.00 


13.50 
9.75 










79 50 






81.25 


No 2 hand fired 


10.25 
13.25 

10.95 

8.75 


8.60 
3.50 

1.30 

7.00 


.50 
.05 

3.00 

3.25 



0.25 

.70 

.40 





3.23 

1.00 


80.65 
82.95 

80.82 

79.60 








No S hand fired 














No. 4, fire under caustic 
pot, hand fired 

No. 5, split bridge, hand 
fired 


























No. 6, with smoke-pre- 
vention device 


7.25 
7.15 
8.15 


12.00 
12.15 
11.10 

















80.75 


No. 7, with smoke-pre- 














80.70 


No. 8, with smoke-pre- 














80.75 



















TABLE 16. 

RELATION OF CO AND COMBUSTION-CHAMBER TEMPERATURES. 

(U. S. Geological Survey). 



Per Cent of Black Smoke. 



to 10 



10 to 20 



20 to 30 



30 to 40 



40 to 50 



50 to 60 



Number of tests 

Average per cent of smoke 

Average per cent of CO in flue gases 

Average per cent unaccounted for in heat 
balance 



37 

0.05 



9.14 



18 
7.1 
0.11 

10.60 



56 
15.5 
0.11 



51 
24.7 
0.14 

10.93 



3fi 
34.7 
0.21 



17 
43.1 
0.33 

13.41 



4 

52.9 
0.35 

13.34 



Number of tests * 

Average combustion-chamber temperature 
(°F.) 



26 
2180 



2215 



48 
2357 



45 
2415 



32 

2450 



17 
2465 



4 

2617 



Temperatures in combustion chamber were not determined on all tests. 



as will be seen from Table 17. Carbon monoxide is a colorless gas 
and its presence in the chimney gases cannot be detected by the fire- 
man, consequently the absence of smoke is not an infallible guide for 
perfect combustion. Since the heat of combustion of C to CO is but 
4380 B.t.u. against 14,540 B.t.u. for complete combustion of C to CO2 
this loss may be expressed 



C 



- C 



(14,540 - 4380) CO 

CO2 + CO 
10,160 CO 



(22) 



CO2 + CO' 



64 



STEAM POWER PLANT ENGINEERING 



in which 

/i2 = the loss in B.t.u. per pound of fuel. 

C = proportional part of carbon in the fuel which is burned and 
passes up the stack. 

CO2 and CO are percentages by volume. 

This loss may be reduced to a negligible quantity in a properly de- 
signed and carefully operated furnace. In fact the loss from this cause 
is often exaggerated and seldom exceeds 1 per cent of the total heat 
value of the fuel except during the few moments following the replen- 
ishing of a burned-down fire with fresh fuel or when the supply of air 

TABLE 17. 
LOSS DUE TO INCOMPLETE COMBUSTION OF CARBON TO CARBON 

MONOXIDE. 



3 

I 



% 

1 
a 

8 

c 

6 

1 




Per Cent of CO2 in the Flue Gas by Volume. 


G 


8 


10 


12 


14 


IG 


0.2 


328 

2.2 


248 

1.7 


199 

1.3 


168 

1.1 


144 

1 


126 

0.8 


0.4 


635 

4.3 


484 

3.3 


390 

2.6 


327 

2.2 


282 
1.9 


248 

1.7 


0.6 


925 

6.3 


709 

4.8 


575 

3.9 


474 

3.2 


417 

2.8 


367 

2.5 


0.8 


1192 

8.1 


923 

6.3 


750 

5.1 


635 

4.3 


549 

3.7 


495 

3.4 


1.0 


1494 

10.2 


1128 

7.7 


923 

6.3 


780 
5.3 


676 

4.6 


596 

4.1 


1.2 


1690 

11.5 


1321 

9 


1085 

7.4 


923 

6.3 


801 

5.4 


708 
4.8 


1.4 


1920 

13.1 


1512 

10.3 


1248 

8.5 


1061 

7.2 


924 

6.3 


819 

5.6 


1.6 


2104 

14.3 


1693 

11.5 


1400 

9.5 


1193 

8.1 


1040 

7.1 


924 

6 3 


1.8 


2340 

16 


1865 

12.7 


1549 

10.5 


1321 

9.0 


1151 

7.8 


1025 

7 


2.0 


2537 

17.2 


2030 

13.8 


1690 

11.5 


1450 

9.9 


1270 

8.6 


1129 

7.7 



Large type gives the loss in B.t.u. 
per cent loss, assuming a calorfic value 



per pound of carbon. Small type gives 
of 14,540 B.t.u. per pound of carbon. 



the 



FUELS AND COMBUSTION 65 

is checked to meet a sudden reduction in load. In improperly designed 
furnaces in which the volatile gases are brought into contact with the 
cooler boiler surface before combustion is complete, the carbon monoxide 
may be reduced in temperature below its ignition point and conse- 
quently will fail to combine with the oxygen. In such a case the loss 
may prove to be a serious one. 

High efficiencies necessitate minimum air excess, hence the presence 
of a small amount of CO may be expected in the flue gas. In a number 
of recent tests of modern central station boilers operating at 150 to 250 
per cent of standard rating, the loss due to the escape of CO in the 
flue gas ranged from 0.2 to 1.95 per cent of the heat value of the fuel 
(Western bituminous) with a general average, extended over several 
days, of 0.4 per cent. In these tests the per cent of CO2 in the flue gas 
ranged from 11.95 to 15.45. The CO content appears to increase with 
the increase in CO2 and furnace temperature as shown in Fig. 10, the 
curves of which are based on tests of a 250 horsepower Heine boiler, 
hand fired. (Journal Western Society of Engineers, June 1907, p. 285.) 
Almost complete absence of CO is to be expected with large air excess 
in any well designed furnace, but it is possible for a high percentage 
of CO and a great excess of air supply to exist at the same time, though 
this combination is not Jikely to occur in a properly designed furnace 
except at very low rates of combustion. 

38. Loss of Fuel Through Grate. — The refuse from a fuel is that 
portion which falls into the pit in the form of ashes, unburned or par- 
tially burned fuel and cinders. 

In steam boiler practice the unconsumed carbon in the ash pit ranges 
from 15 to 50 per cent of the total weight of dry refuse depending upon 
the size and quality of coal, type of grate and rate of driving. The 
loss resulting from this waste of fuel ranges from 1.5 to 10 per cent or 
more, of the heat value of the fuel. It is impossible to assign a minimum 
value because of the various influencing factors, but numerous tests 
of recent installations, equipped with mechanical stokers, indicate that 
actual loss ranges from 1.5 to 5 per cent of the heat value of the fuel 
at normal driving rates. Coal which necessitates frequent slicing is 
apt to give greater losses from this cause than a free burning coal. 

Extensive tests conducted by the American Gas & Electric Company, 
(Reginald Trautschold, Power, Feb. 22, 1916, p. 256) show that the 
actual yearly loss due to combustible in the refuse is not directl}' pro- 
portional to the combustible content but increases as shown by the 
"actual loss" curve in Fig. 11. Thus, the reduction of the combustible 
content from 10 per cent to 5 per cent effects a yearly saving in the 
ratio of 12.98 to 5.83 instead of 10 to 5. 



66 



STEAM POWER PLANT ENGINEERING 



In traveling grates in which a large percentage of the fine fuel falls 
through the front end of the grate a special hopper is ordinarily in- 
stalled in the ash pit which reclaims most of it. (See Fig. 130.) 
If he = calorific value of combustible in the dry refuse, 
y = percentage of combustible in the dry refuse, 
a = percentage of ash in the coal as fired, 
hs = heat loss in the refuse, B.t.u. per pound of coal as fired, 

' 100 



UOO -y) 



(23) 



For the average boiler test the calorific value of the combustible in the 
dry refuse may be taken as that of pure carbon or of the combustible in 
the coal (see end of paragraph 20) but for accurate results calorimetric 
determinations are necessary. 



10 
















~ 




































7 














- 


^ 


T 


12 


















































/ 


















y 




















































/ 
















y 
























































/ 








































































/ 












^. 


/ 


























































/ 










1^. 






















clU 






































.^ 


/ 










, 

-^^ 


r 


r 




















f' 




































^V 










} 


r 






















































, 


.^ 


7 






/* 






























> 
































.<-y 






V 


; 
























































Vo^/ 




f 


y 


































5, 




























</ 




• 


^ 


, 






























































/ 






k* 








































o 6 


























/ 




/ 


































































/ 




Y 














































^5 






















/ 




y 


































































/ 


/ 




















































|, 


















/^ 






































































/ 


/* 
























































^ 3 














/ 


/* 




































































^ 




























































9 










/, 


4^ 






































































// 


































































1 






^ 






































































/ 









































































/ 
















1 


























































I 


£ 


I 


lii 




ei 


It 


€( 


:>a 


II 


3 
0^ 


33 




ae 


3 

tc 


c 

c 


01 


1 





1 
9tJ 


1 


1 
ej 


2 


1 

A 


3 


1 
P 


4 


i 
C 


5 
ei 


i 
It 


6 


1 


7 


IS 



Fig. 11. Coal Loss Due to Combustible in Ash. 



: 2«. Superheating the Moisture in the Air. — The loss due to this 
cause is a minor one, though on hot, humid days it may be appreciable. 
This loss may be expressed 

/i4 = Mc {tc - t), (24) 

in which 

hi = B.t.u. lost per pound of fuel, 

M = weight of moisture introduced with the air per pound of fuel, 
c = mean specific heat of water vapor, t to U deg. fahr., 
t = temperature of air entering the furnace, deg. fahr., 
tc = temperature of chimney gases, deg. fahr. 

~ " M = zwvA, (25) 



t 



FUELS AND COMBUSTION 67 

in which 

z = relative humidity (see paragraph 470), 

w = weight of 1 cubic foot of water vapor at t deg. fahr. (this may 

be taken directly from steam tables), 
V = volume of 1 pound of dry air at t deg. fahr., cubic feet, 

A = weight of dry air supplied per pound of fuel burned. 

30. Loss Due to Moisture In the Fuel. — Moisture in the fuel repre- 
sents an appreciable loss in economy if present in large quantities, since 
the heat necessary to evaporate it into superheated steam at chimney 
temperature is lost. Firemen occasionally wet the coal to assist 
coking or to reduce the dust, but moisture thus added necessarily 
reduces the theoretical furnace efficiency. Under certain conditions 
wet coal may give a higher evaporation than dry coal, that is, the 
moisture may assist in packing the fuel and thus reduce loss through the 
grate, and in case of thin fires, reduce air excess. The action of the 
ijioisture is purely mechanical. (See paragraph 99.) 

The loss due to evaporating the moisture may be expressed 
in which h=Wl\-c,{t- 32) + c' (t. - 1%^ (26) 

h = B.t.u. lost per pound of fuel. 
W = weight of free moisture per pound of fuel. 
X = total heat of one pound of saturated steam above 32 deg. fahr., 
corresponding to the temperature at which evaporation takes 
place. 
Ci = mean specific heat of water, 32 to t deg. fahr. 
t = temperature of the fuel, deg. fahr. 
c' = mean specific heat of the water vapor, tc to t deg. fahr. 
tc = temperature of the chimney gas. 

t' = temperature, at which evaporation takes place, deg. fahr. 
The temperature at which evaporation begins is low because of 
the low partial pressure of the vapor in the gaseous products of 
combustion and may range from 70 to 120 degrees, depending 
upon the composition of the gases and the amount of moisture 
evaporated. Fortunately, the term X — c't' is practically con- 
stant for a wide range of t' and consequently a knowledge of 
the actual value for each set of conditions is not necessary. 
Assuming c' = 0.46 ^nd taking X from the steam tables for all values 
ranging from f = 70 to f = 120 degrees, we find that X — 0.46 t' = 
1058.6. Substituting this value in equation (26) and reducing, 

h = w [1090.6 - t + 0.46 Q. (27) 

* For all engineering purposes Ci may be taken as unity and in the following equa- 
tions it has been considered as such. 



68 STEAM POWER PLANT ENGINEERING 

31. Loss Due to the Presence of Hydrogen in the Fuel. — The hydro- 
gen in any fuel which is not rendered inert by oxygen burns to water 
and in so doing liberates 62,000 B.t.u. per pound. All of this heat is 
not available for producing steam in the boiler, since the water formed 
by combustion is discharged with the flue gases as superheated steam 
at chimney temperature. This loss is equal to 

he = 9H (1090.6 - t + 0.46 to), (28) 

in which 

he = B.t.u. lost per pound of fuel, 

H = weight of hydrogen per pound of fuel burned. 

All other notations as in equations (26) and (27). 

With anthracite coal this loss is approximately 2.5 per cent of the 
total heat value of the combustible and with bituminous coal it runs as 
high as 4.5 per cent. 

33. Loss Due to Visible Smoke. — Visible smoke consists of carbon in 
a flocculent state and ash mixed with the products of combustion. It is 
seldom evident in connection with anthracite coal and is generally 
associated with bituminous fuel. A smoky chimney does not neces- 
sarily indicate an inefficient furnace, since the losses due to visible 
smoke generation seldom exceed one per cent; * as a matter of fact, a 
smoky chimney ma}^ be much more economical than one which is 
smokeless. That is to say, a furnace operating with minimum air sup- 
ply may cause dense clouds of smoke and still give a higher evaporation 
than one made smokeless by a very large excess of air. There will be 
some loss due to carbon monoxide, unburned hydrocarbons and soot 
in the former case, but this may be more than offset by the excessive 
losses caused by the heat carried away in the chimney gases in the 
latter. The amount of combustible in the soot and cinders deposited 
on the tubes and in various parts of the setting seldom exceeds one per 
cent of the calorific value of the fuel. 

Smoke has become such a public nuisance, particularly in the larger 
cities, that special ordinances prohibiting its production have been 
enacted and violators are subject to heavy fines. Effective enforce- 
ment of these ordinances renders smoke production very costly and 
the problem of smokeless combustion becomes a momentous one. 

The subject of smoke prevention and smoke-prevention devices is 
discussed at some length in Chapter V. 

33. Radiation and Unaccounted For. — These losses are usually de- 
termined by difference. That is, the difference between the heat 
represented in the steam and the losses just mentioned are charged to 

* See paragraph 92. 



FUELS AND COMBUSTION 69 

"unconsumed hydrogen and liydrocarbons, to radiation and unaccounted 
for." Unless accurate observations have been made in determining 
the various factors entering into the heat balance the radiation and 
unaccounted for loss may represent a large percentage of the total 
heating value of the coal. Careful tests on well-designed boiler furnaces 
show that the radiation loss seldom exceeds two per cent. In case of 
very poorly installed settings or when the rate of driving is very low the 
radiation loss may be considerably more than this. An examination 
of the data from carefull}' conducted tests of modern boiler furnaces 
will show that the ''radiation and unaccounted for" items range from 
2 to 6 per cent with an average of about 4 per cent. Soot deposited 
on the boiler tubes and throughout the setting, and cinders blown 
out the stack under high draft pressures may greatly increase the un- 
accounted for loss, unless means are available for determining these 
factors. For data pertaining to the loss represented by visible smoke, 
soot, and cinders, see paragraph 92. 

34. Heat Balance. — Any chart giving the distribution of the various 
heat items constitutes a heat balance. The greater the number of 
subdivisions the more readily is it possible to locate the source of loss. 
The various factors entering into the commercial boiler heat balance 
as recommended by the American Society of Mechanical Engineers 
are itemized in Table 18.* According to this code the heat distribution 
is expressed in terms of ''dry coal" or "combustible." When com- 
paring the performance of different installations this offers a most 
satisfactory basis, but the operating engineer in tracing out the source of 
heat loss with a view of bettering operation is chiefly concerned with 
"coal as fired" and for this reason the heat balance is commonly ex- 
pressed in terms of the latter. It is impracticable to assign specific 
limiting values to a general heat balance because of the wide range in 
the various influencing factors, such as nature and quality of fuel, 
type of furnace and grate, rate of driving and the like, but for a rough 
approximation, Table 18 may be taken as representative practice. 

The heat balance in Table 18 refers to boiler in continuous oper- 
ation and does not include standby losses. (See paragraph 35.) 

The calculations of the various items included in the heat balance 
are best illustrated by a specific example. 

Example 11. Calculate the various heat losses from the following 
data: 

Heat absorbed by the boiler, 76 per cent of the calorific power of the 
coal as fired. 

♦Rules for Conducting Evaporation Tests of Boilers, A.S.M.E., Code of 1916. 



70 



STEAM POWER PLANT ENGINEERING 



Analysis of coal as fired : p^^ ^ent. Per Cent, 

Carbon 65 Ash and sulphur 13 

Oxygen 8 Free Moisture 8 

Hydrogen 5 Nitrogen 1 

Calorific value as fired, 11,850 B.t.u. 

Flue-gas analysis : p^^ ^ent. Per Cent. 

CO2 12.8 CO 0.6 

O2 5.4 N2 81 . 2 (by difference) 

Temperature of air entering furnace, 70 deg. fahr. ; temperature of 
flue gases, 470 deg. fahr.; temperature of the steam in the boiler, 340 
deg. fahr.; relative humidity of air entering furnace, 80 per cent; com- 
bustible in the dry refuse, 20 per cent. 

The heat distribution may be referred to the coal as fired, dry coal 
or combustible. In this problem it is referred to the coal as fired. 



CALCULATION. 

The combustible in the ash referred to the coal as fired is 



20 X 13 



100 - 20 

= 3.25 per cent or 0.0325 lb. per pound of coal. Taking this as carbon 
the actual weight of carbon burned and appearing in the chimney gas 
is 0.65 - 0.0325 = 0.6175 lb. per lb. of coal as fired. 

The weight of dry chimney gas per pound of carbon is equation (11) 

4 X 12.8 -t- 5.4 + 700 _ ,, ,, 
^' ~ 3 (12.8 + 0.6) ' ~ ^^'^^ 

For the carbon actually burned this is 18.82 X 0.6175 = 11.62 lb. 
per lb. of coal as fired. 

The dry air supplied per pound of coal as fired is (equation 12) 

_ 3.032 X 81.2 _ 
^'- 12.8+0.6 -^^'^^' 

For the carbon actually burned this is 18.36 X 0.6175 = 11.34 lb. 
per lb. of coal as fired. 



DISTRIBUTION OF ACTUAL LOSSES PER POUND OF COAL AS FIRED. 



Equa- 
tion. 



(21) 

(22) 

(23) 
(27) 
(28) 

(25) 



Loss. 



Heat absorbed by boiler. 
Dry chimney gas 



Incomplete combustion, . . . 

Combustible in refuse 

Moisture in the fuel 

Moisture from combustion 

of hydrogen 

Moisture in the air 



Radiation and unaccounted 

for 

Total 



Calculation. 



0.76 X 11,850 
11.62 X (470-70)0.24 

0.6175 X 10,160 X!:; ^'^ 



12.8 + 0.6 
0.0325X14,600 
0.08 [1090.6 + 0.46 X 470 - 70] 

9X0.05(1090.6+0.46 X 470 - 70] 
0.08 X 0.00115 X 13.2 X 11.34 
X 0.46 (470 -70) 



By difference. 



B.t.u. 



9,006 
1,115 

280 

474 
99 

556 

25 

295 



11,850 



Per 
Cent. 



76.00 

9.40 

2.36 

4.00 
0.83 

4.70 

0.20 

2.51 



100.00 



FUELS AND COMBUSTION 



71 



TABLE 18. 

TYPICAL HEAT BALANCE. — BITUMINOUS COAL. BASED ON COAL AS FIRED. 



Heat absorbed by the boiler 

Loss due to the evaporation of free moisture in the 

coal 

Loss due to the evaporation of water formed by the 

combustion of hydrogen 

Loss due to heat carried away by the dry flue gas. . 

Loss due to carbon monoxide 

Loss due to combustible in the ash and refuse 

Loss due to heating moisture in the air 

Loss due to unconsumed hydrogen, hydrocarbons, 

radiation and unaccounted for 

Calorfic value of the coal 



Excel- 
lent 
Prac- 
tice. 



Good 
Prac- 
tice. 



Aver- 
age 

Prac- 
tice. 



Poor 
Prac- 
tice. 



Per Cent of Calorific Value of 
Coal as Fired. 



80.0 


75.0 


65.0 


O.o 


0.6 


0.6 


4.2 
10.0 
0.2 
1.5 
0.2 


4.3 
13.0 
0.3 
2.4 
0.2 


4.3 
17.5 
0.5 
4.5 
0.3 


3.4 


4.2 


7.3 


100. 


100.0 


100.0 



60.0 

0.7 

4.4 
20.0 
1.0 
5.5 
0.4 

8.0 
100.0 



35. Standby Losses. — The heat balance as ordinarily calculated refers 
only to the heat distribution for continuous operation over a limited 
period of time. It does not represent average operating conditions 
since the various standby losses are not considered. These include: 
(1) heat lost in shutting down boilers; (2) coal required to start up cold 
boilers; (3) coal burned in banking fires, and (4) heat discharged to 
waste in ''blowing off" and in cleaning boilers. The magnitude of the 
standby losses depends upon the size and character of the boiler equip- 
ment and the conditions of operation and may range from 5 to 15 per 
cent or more of total heat generated (yearly basis) . Thus, a continuous 
24-hour full load test may show that 80 per cent of the heat of the coal 
is absorbed by the boiler, but when the heat represented by a month's 
evaporation is divided by the heat of the fuel fed to the furnace during 
the same period, the efficiency may drop to 70 per cent or lower. The 
standby losses are dependent upon so may variable factors that even 
average figures may be misleading unless limited to a narrow field of 
operation. The data in Table 19 compiled from carefully conducted 
tests at the central heating and power plant of the Armour Institute of 
Technology, serve to illustrate the extent and influence of the standby 
losses on the overall efficiency in a specific case. 

Table 20 gives the weight of coal burned in shutting down boilers, 
starting up cold boilers and in banking fires for a number of Chicago 
plants. 



72 



STEAM POWER PLANT ENGINEERING 



TABLE 19. 

INFLUENCE OF STANDBY LOSSES ON OVERALL BOILER AND FURNACE 

EFFICIENCY. 



Period Covered by Test. 



Number of hours in month 

Hours in service 

Hours banked, or out of service 

Per cent of rating developed, average for 

month 

Total water: 

Fed to boiler, pounds 

" Blowing off," pounds 

Net evaporation 

Total coal: 

Fed to furnace, pounds 

Burned in banking, etc., pounds 

Used for evaporation, pounds 

Apparent evaporation per pound of coal 

fed to furnace, pounds 

Actual evaporation per pound of coal 

used for evaporation, pounds 

Gross overall efficiency of boiler and 

furnace, per cent 

Overall efficiency, deducting standby 

losses, per cent 



January. 



744 

708 

36 

133.0 

11,375,390 

74,800 

11,366,340 

1,360,370 

3,680 

1,356,690 

8.35 

8.38 

71.9 

72.0 



October. 



744 
624 
120 

60.2 

5,235,420 

39,870 

5,230,210 

728,360 

13,850 

714,510 

7.19 

7.32 

61.8 

63.2 



July.t 



744 
153 
591 

13.2 

791,610 

16,150 

789,990 

158,960 

37,610 

121,350 

4.98 

6.51 

44.0 

57.6 



* January and October tests: 350 horsepower Stirling boiler equipped with chain grate, feed water 
205 deg. fahr., pressure 100 pounds gauge, Illinois No. 3 washed nut. 
t July test: 250 horsepower ditto. 



TABLE 20. 

COAL BURNED DURING BANKING PERIODS. 













Coal Fed to 








Ratio 






Furnace, Lb 




Rated 
Capacity 
of Boiler. 


Kind of Stoker. 


Heat- 
ing to 
Grate 


Kind of Coal. 


Hours 
Banked. 


per Boiler 
Hp.-hr. 


c 












A 


B 




250 


Stationary grate 


35 


Buckwheat 


8 


0.20 


0.35 




500 


Chain grate 


65 


Bit. scrg. 


13 


0.40 


0.52 


1000 


350 


Chain grate 


40 


Bit. No. 3 


9 


0.32 


0.62 


1600 


250 


Chain grate 


48 


Bit. scrg. 


7 


0.35 


0.71 


1450 


1200 


Underfeed 


82 


Bit. scrg. 


10 


0.18 


0.20 


2600 


550 


Underfeed 


66 


Bit. scrg. 


9 


0.29 


0.37 


1165 


150 


Stationary grate 


40 


Bit. mine run 


12 


0.58 


0.69 


560 


75 


Stationary grate 


48 


Poc. lump 


12 


0.81 


0.95 


300 


400 


Murphy 


52 


Bit. scrg. 


13 


0.26 


0.33 


1350 



{A) Coal fired during banking period. 

(B) Coal fed to furnace during banking period including that required to put boiler into service at end 
of banking period. 

(C) Coal fed to furnace to put cold boiler into service, lb. 

* These values are for specific cases only. The range in practice is so wide that average values are 
misleading. 



I 



FUELS AND COMBUSTION 73 

The loss due to ''blowing off" depends largely upon the quality of 
the feed water. Water containing considerable scale forming element 
requires frequent blowing off, the amount discharged varying from one 
half to two gauges of water. For example, the 350 horsepower Stirling 
boiler in the power plant of the Armour Institute of Technology (Table 
19, Col. 1) is blown off once in 24 hours when in continuous operation, 
the amount averaging 3 inches as indicated by the water gauge. For 
one month this totals 74,800 pounds. The heat lost is 74,800 (338 - 
205) = 9,200,000 B.t.u., approximately, or sufficient to evaporate 
9050 pounds of water from a feed temperature of 205 deg. fahr. to steam 
at 100 pounds gauge. This amount should be deducted from the 
water fed to the boiler in calculating the net evaporation (the quality 
of the steam, of course, being taken into consideration). Compared 
with the monthly evaporation this loss is neghgible, though it repre- 
sents an appreciable loss per se. 

The steam required in blowing soot from the tubes of a return tubular 
boiler ranges from 250 to 400 pounds of steam per cleaning with ''hand 
blowing" and from 200 to 350 pounds with mechanically operated 
"soot blowers." For water tube boilers the range is considerably 
greater, depending upon the size of the units and the time interval 
between cleanings. A rough approximation is 500 to 750 pounds for 
hand blowing and 400 to 600 pounds for mechanical blowers incorpo- 
rated within the setting. 

Tests of Hand and Mechanical Soot Blowers, Power, July 13, 1915, p. 48. 

36. Inherent Losses. — The heat balance as ordinarily calculated gives 
the distribution of the actual losses. Some of these losses may be con- 
siderably reduced or even entirely eliminated, while others are in- 
herent and cannot be prevented. A heat balance gi\ing the extent of 
the inherent losses will show at a glance where improvement may be 
made and where further gain is impossible. A boiler and furnace may 
be perfect in operation and still fail to utilize the total heat value of 
the fuel. For example, in the modern boiler (without an economizer 
or its equivalent) the flue gas cannot be lowered below the temperature 
of the heating surface with which it was last in contact. Since this 
temperature corresponds to that of the steam in the boiler, we have as 
the inherent losses: 

1. Heat absorbed by the theoretical weight of dry chimney gases 
in being heated from boiler room to boiler steam temperature. 

2. Heat required to evaporate and superheat the moisture in the fuel 
from boiler room to boiler steam temperature. 

3. Heat required to evaporate and superheat the H2O formed by 



74 



STEAM POWER PLANT ENGINEERING 



the combustion of hydrogen in the fuel from boiler room to boiler steam 
temperature. 

4. Heat required to superheat the moisture in the air (theoretical 
requirements) from boiler room to boiler steam temperature. 

Example 12. Determine the inherent losses from the data given in 
Example 11. 

DISTRIBUTION OF INHERENT HEAT LOSSES PER POUND OF COAL AS FIRED. 



1. Inherent loss in the dry chimney gas, 

9.26* X (340 -70)0.24 

2. Inherent loss due to moisture in coal, 

0.08 (1090.6 - 70 + 0.46 X 340) 

3. Inherent loss due to H2O formed by the combustion of 

hydrogen, 9 X 0.05 (1090.6 - 70 + 0.46 X 340) 

4. Inherent loss due to " humidity " of the air, 

0.8 X 0.00115 X 13.3 X 892 * X 0.46 (340 - 70) 

5. Heat absorbed by ideal boiler (by difference) 



* See example 7, paragraph 22. 

A comparison of the actual and inherent losses in percentages of coal 
as fired is as follows: 




1. Dry chimney gases 

2. Incomplete combustion 

3. Combustible in the refuse 

4. Moisture in the air 

5. Moisture in the coal 

6. Moisture due to combustion of hydrogen 

7. Radiation and unaccounted for 

8. Heat absorbed by the boiler 



100.00 



Actual. 


Inhere at. 


9.40 


5.06 


2.36 


0.00 


4.00 


0.00 


0.20 


0.11 


0.83 


0.79 


4.70 


4.47 . 


2.51 


0.00 


76.00 


89.57 



100.00 



The difference between the actual and inherent loss is designated 
as preventable. Although the losses due to "incomplete combustion/' 
"combustible in the refuse," and ''radiation and unaccounted for" 
are theoretically preventable it is almost impossible to entirely ehminate 
them in practice. The minimum practical loss depends upon the nature 
of the equipment, grade of fuel and rate of driving, and must be deter- 
mined for each installation by actual test. This is also true for the 
"preventable" loss in the dry chimney gases and that due to the mois- 
ture in the air, moisture in the coal and moisture resulting from the 
combustion of hydrogen, 



FUELS AND COMBUSTION 75 

Since the ideal or perfect boiler under the specified conditions is able 
to absorb only 89.57 per cent of the calorific value of the coal it is evident 
that the actual boiler has a true efficiency of 76 -r- 0.8957 = 84.8 per 
cent. 

If an economizer is used the inherent losses become less since the flue 
gas may be reduced to a temperature considerably lower than that of 
the steam but they can never be entirely ehminated unless the flue gas 
is discharged at the same temperature as that of the air entering the fur- 
nace. 

37. Selection and Purchase of Coal. — Perhaps no single item in the 
operation of an existing plant or in the design of a new plant affords 
such an opportunity for effecting economy as the selection of fuel. Care- 
ful investigations have shown that almost any fuel can be efficiently 
burned in suitably designed special furnaces so that the problem of 
selecting a fuel for a proposed installation requires experience with the 
different kinds of equipment in addition to a thorough knowledge 
of the characteristics of various fuels. For existing plants the prob- 
lem is largely a matter of testing. In many cases it has been found 
advisable to redesign furnaces to utilize a low-grade fuel rather than 
purchase an expensive coal. The following information is useful in 
deciding on the coal best adapted for a plant: * 

a. Type and size of boilers and furnaces. 

h. Load conditions, average and maximum loads. 

c. Draft available and method of control. 

d. Character of coals offered or available. 

1. Moisture and its effect on weight of combustible. 

2. Volatile matter and its relation to type of furnace. 

3. Ash ; its amount and its f usibiHty and tendency to clinker. 

4. Sulphur; the amounts and how combined. 

5. Heating value, calorimeter determination. 

6. Coking quahties of the coal. 

7. Storage and tendency to spontaneous combustion. 

e. Relation of the size of coal to the equipment. 

After the desired grade of fuel has been decided upon the next step 
is to enter into an agreement with the dealer whereby the delivery of 
that particular fuel may be depended upon. The important items to be 
considered in the specifications are: 

a. A statement of the amount and character of the coal desired. 
6. Conditions for delivery. 

* The Purchase of Coal, Dwight, T. Randall. Jour. A.S.M.E., Sept. 1911, p. 987. 



76 



STEAM POWER PLANT ENGINEERING 



c. Disposition to be made of the coal in case it is outside the Hmits 
specified. 

d. Correction in price for variation in heating value and in moisture 
and ash content. 

e. Method of sampling.* 

/. By whom analyses are to be made. 

In specifying the character of the coal desired for the average small 
plant every essential requirement of the purchaser may be fulfilled by 
confining them to the four following characteristics: 

Moisture. 

Ash. 

Size of coal. 

Calorific value of the coal. 

Although moisture is a great and uncertain variable, and the producer 
can exercise no control over this factor, still the purchaser should pro- 
tect himself against excessive 
moisture by stipulating an 
amount consistent with the 
average inherent moisture in 
the coal, and proper penalty 
should be fixed for delivery in 
excess of the amount allowed, 
a corresponding bonus be- 
ing paid for delivery of less 
than contract amount. Con- 
siderable attention should be 
given to the percentage of 
earthy matter contained. The 
amount of earthy matter 
usually fixes the heating value 
of the coal, since the heating 
value of the combustible is 
practically constant. The ef- 

Per Cent of Ash in Dry Coal f^^^ ^^ ^gj^ ^^ ^-^Q heat ValuC 

Influence of Ash on Fuel Value ^f Illinois screenings as fired 

of Dry Coal. ^^^^^ a B. & W. boiler with 

chain grate is shown in Fig. 12. This value varies with the different 

types of boilers, grates, and furnaces, but is substantially as illustrated. 

The amount of refuse in the ash pit is always in excess of the earthy 

* See Coal Sampling Methods, Report of Committee on Prime Movers, Trans. 
National Electric Light Association, N. Y. City, 1916. 



lUU 






\ 




























90 








^ 


\ 


































\ 






















80 














\ 


S 
































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70 


















s 


\ 
































N 














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V 


































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5'" 


























\ 








30 




























\ 
































\ 






20 


— 


Influeiice of Ash on Fuel Value of Dry 
Coal,(Illinois Screenings) 
B. & W. Boiler, Chain Grate. 
Screenings with 12.5 Per Cent Ash 
taken at 100. 






\ 












\ 




10 


• 






\ 






















1 








\ 




















Jour 

1 


a.y.E. 


Oct. 1906 

1 


p4M 





Fig. 12. 



I 



FUELS AND COMBUSTION 77 

matter as reported by analysis, except where the amount carried be- 
yond the bridge wall is very large. 

The maximum allowable amount of sulphur is sometimes specified, 
since some grades of coal high in sulphur cause considerable clinkering. 
But sulphur is not always an indication of a clinker-producing ash, 
and a more rational procedure would be to classify a coal as clinker- 
ing or non-clinkering according to its behavior in the particular furnace 
in question, irrespective of the amount of sulphur present. An analysis 
of the various constituents of the ash is necessary to determine whether 
or not the sulphur unites with them to produce a fusible slag, and as 
such analyses are usually out of the question on account of the expense 
attached they may well be omitted. Ash fuses between 2300 and 
2600 deg. fahr. and if the formation of objectionable clinker is to be 
avoided the furnace must be operated at temperatures below the fusing 
temperature. Several large concerns insert an ''ash fusibihty" clause 
in their coal specifications. For a description of ash fusion methods as 
practiced by various concerns consult Transactions of the National 
Electric Light Association, Report of the Committee on Prime Movers, 
1916. 

The heating value of the coal as determined by a sample burned in an 
atmosphere of oxygen does not give its commercial evaporative power, 
since this depends largely upon the composition of the fuel, character 
of grate, and conditions of operation. It serves, however, as a basis 
upon which to determine the efficiency of the furnace. In large plants 
where a number of grades of fuel are available it is customary" to con- 
duct a series of tests with the different grades and sizes, and the one 
which evaporates the most water for a given sum of money, other con- 
ditions permitting, is the one usually contracted for. In designing a 
new plant particular attention should bie paid to the performance of 
similar plants already in operation, and that fuel and stoker should be 
selected which are found to give the best returns for the money. Where 
smoke prevention is a necessity the smoke factor greatly influences the 
choice of fuel and stoker. 

The Purchase of Coal: Eng. Mag., Mar., 1911; Jour. A.S.M.E., Mar., 1911; 
Power, Apr. 6, 1909, p. 642. 

The Purchase of Coal by the Government under Specifications: Bureau of Mines, 
Bull. No. 11, 1910; U. S. Geol. Survey, Bulletins No. 339, 1908; No. 378, 1909. 

The Fusing Temperature of Coal Ash: Power, Nov. 28, 1911, p. 802. 

The Clinkering of Coal: Trans. A.S.M.E., Vol. 36, 1914, p. 801. 

Jour. A.S.M.E., April 1915, p. 205. Power, Oct. 24, 1916, p. 591. 

fc. 

38. Size of Coal — Bituminous. — Coal is usually marketed in dif- 
ferent sizes, ranging from lump coal to screenings. The latter furnish 



78 STEAM POWER PLANT ENGINEERING 

by far the greater part of the stoker fuel used. The sizes and grades of 
bituminous and semi-bituminous coals vary so much, according to 
kind and locality, that there are no standards of size for these coals 
which are generally recognized. According to the A.S.M.E. Boiler 
Code (1915): 

Bituminous coals in the Eastern States may be graded and sized as 
follows : 

a. Run of mine coal; the unscreened coal taken from the mine 
after the impurities which can be practicably separated have been 
removed. 

b. Lump coal; that which passes over a bar-screen with openings 1} 
inch wide. 

c. Nut coal; that which passes through a bar-screen with Ij inch 
openings and over one with f inch openings. 

d. Slack coal; that which passes through a bar-screen with f inch 
openings. 

Bituminous coals in the Western States may be graded and sized as 
follows : 

e. Run of mine coal; the unscreened coal taken from the mine. 

/. Lump coal; divided into 6-inch, 3-inch, and If inch lump, accord- 
ing to the diamter of the circular openings over which the respective 
grades pass; also 6 by 3 lump and 3 by IJ lump, according as the coal 
passes through a circular opening having the diameter of the larger 
figure and over one of the smaller diameter. 

g. Nut coal; divided in 3-inch steam nut, which passes through an 
opening 3-inch diameter and over IJ inch; IJ inch nut, which passes 
through a IJ inch diameter opening and over a f inch diameter opening; 
and f inch nut, which passes through a f inch diameter opening and over 
a f inch diameter opening. 

h. Screenings; that which passes through a IJ inch diameter opening. 

For maximum efficiency coal should be uniform in size. With 
hand-fired furnaces there is usually no limit to its fineness and larger 
sizes can be used than with stokers. As a rule the percentage of ash 
increases as the size of coal decreases. This is due to the fact that all 
of the fine foreign matter separated from larger coal, or which comes 
from the roof or the floor of the mine, naturally finds its way into 
the smaller coal. The size best adapted for a given case is dependent 
upon the intensity of draft, kind of stoker or grate, and the method 
of firing, and its proper selection often affords an opportunity to effect 
considerable economy. The influence of the size of screenings on the 
capacity and efficiency of a boiler in a specific case is illustrated in Fig. 
13. The curves are plotted from a series of tests conducted with 
Illinois screenings on a 500-horsepower B. & W. boiler, equipped with 
chain grates, at the power house of the Commonwealth Edison Company. 



FUELS AND COMBUSTION 



79 





































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800 








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\ 














80 




/ 


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Influence of Size of Coal on the Capacity 
and Efficiency of a B.& W.Boiler, Cbain Grate 
Heating Surface 5000 Sq.Ft. 
Superheating Surface 1000 Sq.Ft. 






> 


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30 1 

^! 

200 20 1 

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lial 


nche 


s 


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Fig. 13. Influence of Size of Coal on Boiler Capacity and Efficiency. 

Influence of Thickness of Fire. — See paragraph 82. 

Size of Coal: Some Characteristics of Coal as affecting Performance with Steam 
Boilers: Jour. West. Soc. Engrs., Oct., 1906, p. 528. 

39. Washed Coal. — Coal is washed for the purpose of separating from 
it such impurities as slate, sulphur, bone coal, and ash. All of these 
impurities show themselves in the ash when the coal is burned. Screen- 
ings contain anywhere from 5 per cent to 25 per cent of ash and from 
1 per cent to 4 per cent of sulphur. Washing eliminates about 50 per 
cent of the ash and some of the sulphur. Table 21 gives. some idea of 
the effects of washing upon a number of grades of coal. The evapora- 
tive power of the combustible is practically unaffected by washing and 
the greater part of the water taken up by the coal is removed by thor- 
ough drainage. Many coals otherwise worthless as steam coals are 



80 



STEAM POWER PLANT ENGINEERING 



TABLE 21. 

EFFECT OF WASHING ON BITUMINOUS COALS. 
(Journal W.S.E., December, 1901.) 



Before Washing. 
(Percent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



After Washing. 
(Per Cent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



Belt Mountain, Mont 

Wellington Colliery Co., Van- 
couver Island (new coal) . . . , 

Alexandria Coal Co., Crabtree, 
Pa 

De&oto^m.'.'.'.'.'.'.'.'.'.'.'.'.'.'.'.'.'. 

Northwestern Improvement 
Co., Roslyn, Wash 

Luhrig Coal Co., Zaleski, Ohio 

Rocky Ford Coal Co., Red 
Lodge, Mont 

Buckeye Coal and Ry. Co., 
Nelsonville, Ohio 

New Ohio Washed Coal Co., 
Carterville, 111 



18.74 

35.00 

10.60 
18.00 

16.30 
15.80 

25.30 

13.77 

9.48 



3.34 


43.72 




38.00 


1.30 






44.00 


0.57 
1 90 


45.90 




37.80 



1.05 
0.78 



49.04 
55.00 



5.56 

8.90 

6.21 
4.20 

9.70 
8.00 

8.50 

4.30 

4.85 



2.40 



0.61 



0.40 

0.87 



0.89 
0.69 



48.39 
56.90 



57.00 

47.86 
50.90 

47.20 

54.82 

63.00 



rendered marketable by washing, 
as follows : 



Washed coals are usually graded 



Size. 


Screens. 


No. 1 
2 
3 
4 
5 


Over If 
If 


Through 3 
If 

1| 

4 





Numbers 3 and 4 are excellent sizes for use in connection with stokers 
and No. 5 is well adapted to hand furnaces where smoke prevention 
is essential. 

Coal Washing in Illinois. Univ. 111. Bull. No. 9, Oct. 27, 1913. 



40. Powdered Coal. — The use of powdered coal in the manufacture 
of cement and in other industrial processes is an estabhshed success. 
The steadily increasing cost of oil is stimulating the adoption of powdered 
coal in many situations where fuel oil is now burned. As a fuel for steam 
boiler furnaces, however, powdered coal is still in an experimental stage, 
and although published accounts of the various trials are full of promise 
and apparent accomplishment, but few processes have survived. In 



FUELS AND COMBUSTION 81 

view of the high efficiencies effected with bulk coal in connection with 
mechanical stokers it is not likely that powdered coal will be used 
extensively in very large plants since the advantages incident to the 
combustion of the powdered product are more than offset by the cost 
of pulverizing, drying and handhng. Some of the advantages obtained 
in burning powdered coal are: 

a. Complete combustion and total absence of smoke. The coal in 
the form of dry impalpable dust is induced or forced into the zone 
of combustion where each minute particle is brought into contact with 
the necessary amount of air and complete oxidation is effected with 
minimum air excess. 

h. A cheaper grade of coal may be burned; in fact, some grades of 
coal which are burned with only moderate success in bulk may be 
efficiently consumed in the powdered form. 

c. The fuel supply may be readily controlled to meet the fluctuations 
in load and furnace standby losses may be greatly reduced. 

d. The labor of firing is reduced to a minimum. 

The practical objections which may offset these advantages are: 

1. Cost of preparation. 

2. Explosibility of coal dust. 

3. Storage limitations. 

4. Furnace depreciation. 

5. Disposal of ash and slag. / 

6. Refuse discharged through chimney. 

The various advantages and disadvantages are treated at length 
in the following paragraphs. 

41. Types of Powdered Coal Feeders and Burners. — Powdered coal 
burners may be grouped into two general classes: 

1. The dust-feed burner, in which the coal is supphed in the powdered 
form, and 

2. The self-contained burner, in which the coal is crushed, pulverized, 
and fed to the furnace simultaneously. 

The dust may be fed into the furnace by 

1. Natural draft, 

2. Mechanical means, or by 

3. Forced draft. 

The following outline gives a classification of a few of the best-known 
coal-dust burners : 



» 



82 



STEAM POWER PLANT ENGINEERING 



Natural Draft 



Forced Draft , 



r Natural Draft 
Feed 

Brush Feed 
Blower Feed 



Pinther 
Wegener 

, Schwartzkopff 
Rowe 
General Electric 



Compressed Air Atlas 

Ideal 
Blake 



Paddle Wheel 



Dust Feed 



Self-contained 



Since the natural draft type of feeders are not in evidence in boiler 
practice and are little used in industrial furnaces they will not be con- 
sidered here. For an extended study the reader is referred to the ac- 
companying bibliography. 

All of the successful feeders in current practice are of the forced draft 
type. The Rowe coal-dust feeder, Fig. 14, manufactured by the C. O. 
Bartlett and Snow Company, Cleveland, Ohio, is one of the oldest 
examples of this type and although its application is to be found chiefly 
, . in cement kilns it 

Friction Disk 
J 



Coal Hopper 

A 



F-. 



€ 



Screw 
Conveyor 



D 



Air Discharge 



IS 




Fig. 14. Rowe Coal-dust Feeder. 



has been used with 
some success in 
boiler furnaces. Re- 
ferring to the illus- 
tration, powdered 
coal is fed from 
storage bin to 
hopper A and feed 
wormB. The latter 
forces it down spout 
F directly to the 
delivery tube D where it is caught by the air draft and fed into 
the furnace. The amount of feed depends upon the speed of the feed 
worm which is driven by the friction disk / pressing against the flange 
plate H, This disk is moved in or out by a suitable handle so as to get 
any desired speed. The air is furnished by the fan C the amount being 
controlled by the valve E. 

Figure 15 illustrates a forced draft feeder and burner designed by 
A. S. Mann as applied to a water tube boiler at the Schenectady plant 
of the General Electric Company. This installation has been in con- 
tinuous operation for some time and appears to be a successful com- 
mercial application of coal dust burning. Powdered fuel is fed from 
hopper i/ by a variable-speed motor-driven endless-screw S to down 
spout D where it is picked up by a primary current of air and induced 



FUELS AND COMBUSTION 



83 




Gearing' 



Fig. 15. 



Mann Powdered Coal Feeder 
and Burner. 



into the suction opening of vacuum tee V. A secondary air blast at A 
forces the fuel into burner B whence it is discharged into the furnace. 
The air and fuel are thoroughly mixed in the burner and by controlling 
the fuel supply and the various air inlets any rate of feed may be effected 
without stratification. Auxiliary 
air jets discharging into the fur- 
nace break up the stream issuing 
from the burner and prevent the 
fuel particles from leaving the 
combustion chambers before oxida- 
tion is complete. (See paragraph 
42 for further details of this in- 
stallation.) 

Fig. 17 gives a sectional view of 
the Blake apparatus, and is a typi- 
cal example of a self-contained 
system. It comprises a multistage 
centrifugal pulverizer, coal hopper, 
conveyor and fan mounted on a 
single bedplate. Referring to the 
illustration, coal previously crushed to nut size is fed to the hopper from 
the bottom of which it is conveyed by an endless screw to the first 
stage of the pulverizer. The lumps are thrown out radially by cen- 
trifugal force, due to the rapidly revolving bats, and are reduced to a 
dust by percussion and attrition. The largest chamber contains a fan, 
the function of which is to draw the pulverized material successively 

from one cham- 
ber to another 
and finally de- 
Uver it to the 
discharge spouts. 
The air is drawn 
into the fuel 
chamber with 
the coal through 
passage A , and 
also through 
opening B around 
the shaft. After entering the fan chamber, the mixture of coal dust 
and air receives an additional supply of air through opening C. The 
apparatus may be belt-driven or direct-connected and runs at about 
1200 to 1600 r.p.m., requiring 8 to 12 horsepower for its operation. 



Fuel and Air 
Mixing Paddles 




Fuel Outlet 



Fig. 



16. United Combustion Company's 
Fuel Feeder." 



Pulverized 



84 



STEAM POWER PLANT ENGINEERING 



Experience has demonstrated that as much as 14 per cent of moisture 
in the coal has little effect on the pulverization and burning. Several 
boiler plants equipped with this device gave smokeless combustion 
and high efficiency but faulty furnace design caused the system to be 
abandoned. 



Mixture of Pulverized 
Coal and Air 




^^^^^^^^^^^^^^^^^^^^^^^^^^^^^& 



Fig. 17. Blake Coal-dust Feeder. 

43. Boiler Furnaces for Burning Powdered Coal. — The main diffi- 
culty in the commercial application of powdered coal to boiler furnaces 
appears to He in the correct design of furnace and in the distribution 



FUELS AND COMBUSTION 



85 



of the air supply. Several types of feeders and burners are giving the 
best of satisfaction in cement kilns and in other industrial furnaces, 
but when apphed to steam boilers fail to meet requirements. The 
accumulation of slag and rapid deterioration of the furnace lining is the 
chief cause of failure. In burning bulk coal the mass of incandescent 
fuel stores up a quantity of heat to effect distillation and ignition of 
the volatile matter in the green fuel. Since powdered coal is burned 
in suspension a reverberatory furnace or its equivalent is necessary to 
bring about the same result. A large combustion chamber is necessary 
and the shape of the furnace and path of the flame should be such as 
to insure complete combustion and provide a uniform distribution of 



Steam Drum 



6 Motors and Feeders used 
acro6s face of Boiler 

A' 

This pipe can be any 

length (100 ft. or more) 
May be run underground 
or overhead. 
It need.not be straight. 




l-Burner 
6-Burners used across 

__PToorLine___[!!l^^^''^'- 



^^mM 



Fig. 18. General Arrangement of Powdered Coal Burning Equipment at the 
Schenectady Works of the General Electric Company. 



heat over the boiler heating surface without direct impingement of 
flame. Temperature, velocity, volume and direction of current must 
all be considered since there is perhaps no fuel more sensitive to incorrect 
use than coal dust. Several types of furnaces for burning coal dust 
under steam boilers are described and illustrated in a '^Symposium on 
Powdered Coal," Jour. A.S.M.E., Oct. 1914, but few have survived 
the experimental stage and none appears to have solved the problem, 
although favorable mention is made of the Bettington boiler as com- 
mercially exploited in England. An apparently successful installation 
is that of a 474 horsepower water tube boiler at the Schenectady Works 
of the General Electric Company, Fig. 18 gives a diagrammatic, 
longitudinal section through the boiler and furnace of this installation 
and Fig. 18a illustrates the system of locating burners and air passages. 



86 



STEAM POWER PLANT ENGINEERING 




Six feeders and burners, of the type illustrated in Figs. 15 and 16, are 
attached to the one boiler in order to effect flexibility in operation. 
Several auxiliary air ports distributed throughout the setting are 
directed at various angles against the burning currents thereby insur- 
ing perfect stirring action. 
Combustion is virtually 
complete in eight feet of 
travel even at 200 per 
cent rating. Continuous 
loads of 220 per cent rat- 
ing have been readily main- 
tained and a maximum 
load of 265 per cent rat- 
ing has been carried for a 
short period. No difficulty 
whatever is experienced 
with slag, ash and burnt 
brickwork for loads under 140 per cent rating, but with heavy loads 
particles of slag travel with the gas current and cling to the bottom 
row of tubes. The accumulation of this slag is prevented by '^ blow- 
ing off " with a steam jet. A small amount (2 per cent) of flocculent 



Floor 
/Line 



I 



QQOOQjOI 



Burners 



^^ 



4 Slag Pit 

^//////////////////////.//^ 



I 



Fig. 18a. Diagram of Boiler Front Showing Loca- 
tion of Mann Burner and Air Passages. 




Fig. 19. 



Powdered Coal Furnace as Installed Under a Franklin Boiler at the 
American Locomotive Company's Schenectady Plant. 



ash in the form of a very fine powder is discharged with the chimney 
gases. There is no visible smoke, all soot drops in the gas cham- 
bers before reaching the stack and the slag is drawn out once during 
the day to a concrete pit containing water. (For detailed description 



FUELS AND COMBUSTION 87 

of this installation see "General Electric Review," September and 
October, 1915.) 

Fig. 20 gives the general details of a powdered coal furnace as in- 
stalled under a Franklin boiler at the Schenectady plant of the American 
Locomotive Works. This furnace is reported to have been in continued 
service for 18 months without repairs on the furnace walls. Protection 
against depreciation is effected by a coating of slag as shown in the illus- 
tration. No trouble is experienced from coke or cinder-clogged water 
tubes. (See Journal A.S.M.E., Dec. 1916, p. 1000.) 

A number of locomotives have been recently equipped with pow- 
dered coal burners and are apparently successful in operation. (See 
"Pulverized Fuel for Locomotives," Proceedings N. Y. Railroad Club, 
Feb. 18, 1916.) 

The use of pulverized coal appears to be a commercial success at the 
power plant of the M. K. & T. Shops, Parsons, Kan. For a descrip- 
tion of this installation together with results of the boiler tests, see 
National Engineer, May, 1917, p. 175. 

43. Cost of Preparing Powdered Coal. — The cost of drying and grind- 
ing varies with the size and type of equipment, initial moisture con- 
tent of the coal, degree of fineness required and the quantity treated per 
unit of time. Experience shows that the moisture content should be 
reduced to approximately one per cent or less for efficient grinding where 
screens are used and that 95 per cent of the powder should pass through 
a 100-mesh and 80 to 85 per cent through a 200-mesh screen. Dry coal 
is desired because it can be more intimately mixed with air and fed 
regularly to the furnace. Moist coal will clog the feeding mechanism 
and the screen and tends to pack in the storage bins. Machines that 
depend upon air separation for regulating the fineness of the coal have 
no screens to clog and the moisture content need not be less than 5 
per cent, but the power requirements for the grinding increase with 
the moisture content. The average cost of drjdng and grinding, in- 
cluding maintenance and fixed charges, ranges from 25 to 75 cents per 
ton. In stokers of the ''Blake Pulverizer" type, in which the grind- 
ing, drying and feeding are carried on simultaneously in a self-con- 
tained apparatus, the power consumed varies from 2 to 10 per cent of 
the total power developed by the boiler, depending upon the nature 
of the fuel, efficiency of the driving mechanism and the degree of fine- 
ness of the powdered coal; 5 per cent is a fair average. Powdered coal 
sold in the open market ranges from 50 cents to 90 cents a ton above 
the price of the same coal in bulk. 

44. Storing Powdered Fuel. — Most cities limit the storage of pow- 
dered coal to such a small quantity as to interfere seriously with con- 



I 



88 STEAM POWER PLANT ENGINEERING 

tinuity of operation in case of breakdown to the pulverizing or drying 
apparatus. Spontaneous combustion is likely to occur with moist 
coal and since the dry powdered product is exceedingly hygroscopic it 
is necessary to store it in air-tight bins. Powdered coal in quantity 
should always be kept moving and should never be allowed to stand 
more than a day or two. Coal dust in a suspended state is dangerous 
and may cause a serious explosion, but this danger may be minimized 
by the use of equipment which prevents the leakage of the dust. 

45. Efficiency of Powdered Coal Furnaces. — A comparison of a num- 
ber of tests of hand fired and powdered coal furnaces with different 
types of feeders shows a decided gain in efficiency of the powdered coal 
over the hand fired where the fuel is of a low grade. The gain becomes 

j less marked with fuel of fair quality and disappears entirely with good 
j fuel and properly manipulated automatic stokers. Numerous tests 
j of powdered coal installations are recorded showing boiler and furnace 
' efficiency of 77 to 81 per cent, but these figures are readily equaled 
and have been frequently exceeded with bulk coal firing so that other 
factors than fuel economy must be considered in comparing the com- 
mercial value of the two systems. 

46. Depreciation of Powdered Coal Furnaces. — For complete and effi- 
cient combustion of powdered coal high furnace temperatures are es- 
sential. On account of the high temperatures involved and the slag 
produced from the ash the destruction of the furnace lining is very 
rapid. To withstand the intense heat of combustion brick-work of the 
highest quality is essential since common fire brick are soon reduced to 
a liquid slag. A good quality of fire brick will withstand the heat for 
several months without renewal provided the furnace is properly en- 
closed, otherwise the strain of expansion and contraction due to alter- 
nate heating and cooling will crack the brick. Excellent results have 
been obtained from the use of bricks composed chiefly of the refuse of 
a carborundum slag, but the high cost has prevented their general use. 
Fire brick target walls are not recommended for steam boiler practice 
because of the localization of heat. 

Pulverized Coal for Steam Making: Trans. A.S.M.E., Vol. 36, p. 123-169, 1914. 

Pulverized Coal: General Bibliography; Jour. A.S.M.E., Jan., 1914, p. Iv. 

Some Problems in Burning Powdered Coal: Gen. Elec. Review, Sept. and Oct., 1915. 

Firing Boilers with Pulverized Coal: Power, Feb. 14, 1911. 

Use of Pulverized Coal under Steam Boilers: Prac. Engr. U. S., June 1, 1916, p. 490. 

Tests of Pulverized Fuel: Engr. U. S., Apr. 1, 1904; Power, May, 1904, Feb. 14, 
1911. 

Types of Coal Dust Burners: Engr. U. S., Apr. 1, 1904; Jan. 1, 1903; Power, 
Mar., 1904. 

Burning Low Grade Coal Dust: Power, Sept. 12, 1911, p. 393. 



FUELS AND COMBUSTION 89 

47. Fuel Oil. — The recent developnient of oil wells in the Western and 
Gulf States, with the consequent enormous increase in production, has 
given a marked impulse to the use of crude oil for fuel purposes in steam 
power plants. Where economic and commercial conditions permit, it is 
the most desirable substitute for coal. The total absence of smoke and 
ashes, prompt kindling and extinguishing of fires, extreme rate of combus- 
tion, and ease \vith which it can be handled and controlled are marked ad- 
vantages in favor of fuel oil. The reduction in volume and weight over 
an equivalent quantity of coal for equal heating values and the increase 
in boiler efficiency are factors of no mean importance, particularly in con- 
nection with marine or locomotive work. In stationary work the chief 
objections are the difficulty in securing ample storage capacity and the in- 
creased rate of insurance. An objection sometimes raised against fuel 
oil is the increased depreciation of the setting, but in a well-designed set- 
ting this figure is only nominal and of secondary importance. However, 
in spite of the many advantages presented in the use of fuel oil for power 
plant purposes, the comparatively limited supply prevents its adoption 
as a general fuel and limits its use to the plants most favorably located. 

48. Chemical and Physical Properties of Fuel OU. — Crude oil, as 
pumped at the wells, consists principally of various combinations of 
hydrogen and carbon, together with small amounts of nitrogen, oxygen 
sulphur, water in emulsion and silt. The nitrogen and ox>^gen may be 
classified with the moisture and silt as inert impurities. The moisture 
in oil fuel should not exceed 2 per cent, since it not only acts as an inert 
impurity, but must be converted into steam in the furnace and thus 
still further reduces the heat value per pound. The sulphur, though 
combustible, has a low calorific value and is otherwise undesirable. 
From Table 22 it will be seen that the physical properties of oils from 
different localities in the United States differ widely, while the chemical 
constituents vary but slightly. For example, the oils given in the 
table differ greatly in volatihty, specific gravity, and viscosity, but 
have approximately the same percentages of carbon and hydrogen. 
Taking hydrogen and carbon as the principal constituents it is found 
that oils rich in hydrogen are lighter in weight than those rich in carbon. 
Other things being equal, oils rich in hydrogen have a higher calorific 
value than those rich in carbon, but the heavier oils are usually the 
cheaper. The relation between heating value and specific gravity for 
anhydrous California oil is as shown in Table 23. 

The heat value may be closely approximated by means of the follow- 
ing formula (Jour. Am. Chem. Soc, Oct., 1908): 

B.t.u. = 18,650 + 40 (B - 10), (29) 

in which B = degrees Baume at 60 deg. fahr. 



90 



STEAIM POWER PLANT ENGINEERING 



t^ CO 00 CO 00 
(M lO O ^ t^ 

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^ 00 

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HO 



FUELS AND COMBUSTION 



91 



TABLE 23. 

APPROXIMATE RELATION BETWEEN THE HEATING VALUE AND SPECIFIC GRAVITY. 
(Professor Le Conte, University of California.) 



Degrees, 


Specific 


Weight per 


B.t.u. per 


B.t.u. per 


Degrees, 


Baum6. 


Gravity. 


Barrel. 


Pound. 


Barrel. 


Baum6. 


10 


1.0000 


350.035 


18,280 


6,398,600 


10 


11 


0.9929 


347.55 


18,340 


6,374,100 


11 


12 


0.9859 


345.10 


18,400 


6,349,800 


12 


13 


0.9790 


342.68 


18,460 


6,325,900 


13 


14 


0.9722 


340.30 


18,520 


6,302,400 


14 


15 


0.9655 


337.96 


18,580 


6,279,300 


15 


16 


0.9589 


335.65 


18,640 


6,256,500 


16 


17 


0.9524 


333.37 


18,700 


6,234,000 


17 


18 


0.9459 


331.10 


18,760 


6,211,400 


18 


19 


0.9396 


328.89 


18,820 


6,189,700 


19 


20 


0.9333 


326.69 


18,880 


6,167,900 


20 


21 


0.9272 


324.55 


18,940 


6,147,000 


21 


22 


0.9211 


322.42 


19,000 


6,126,000 


22 


23 


0.9150 


320.28 


19,060 


6,104,500 


23 


24 


0.9091 


318.22 


19,120 


6,084,400 


24 


25 


0.9032 


316.15 


19,180 


6,063,800 


25 



Oil that is to be transported or stored or used for fuel inside of build- 
ings should be of the ''reduced" variety, from which the naphtha and 
higher illuminating products have been distilled. The gravities of 
such distillates vary from 20 to 25 degrees Baume, or close to 0.9 spe- 
cific gravity, and their flash points range from 240 deg. fahr. to 270 
deg. fahr.* One barrel of crude oil contains 42 gallons and weighs from 
310 to 350 pounds, according to the specific gravity. Compared with 
coal, oil occupies about 50 per cent less space and is 35 per cent less in 
weight for equal heat values. The comparative heat values of coal 
and oil are approximately as follows: 



B.t.u. per Pound of 


Pounds of Coal Equal 


Barrels of Oil Equal to 


Coal. 


to 1 Barrel of Oil. 


1 Short Ton of Coal. 


10,000 


620 


3.23 


11,000 


564 


3.55 


12,000 


517 


3.87 


13,000 


477 


4.19 


14,000 


443 


4.52 


15,000 


413 


4.84 



49. Efficiency of BoUers with Fuel Oil. — A coal-burning boiler 
which utilizes 80 per cent of the heat value of the fuel is exceptional — 
77 per cent represents very good practice, and 75 per cent a fair aver- 
age for good practice. The great majority of coal-burning boilers, 
however, operate at efficiencies less than 70 per cent. With oil fuel a 

* For relationship between degrees Baume and specific gravity see paragraph 374. 



92 



STEAM POWER PLANT ENGINEERING 



boiler and furnace efficiency of 75 per cent is quite ordinary and 80 per 
cent not uncommon. This increase in efficiency is partly due to the 
fact that the oil is readily broken up and brought into immediate con- 
tact with the necessary air for combustion and loss due to excessive 
air dilution is correspondingly reduced. 

Table 24 gives the theoretical air requirements for different densities 
of fuel oils and Table 25 the air excess for various efficiencies. These 
tables were compiled by C. R. Weymouth (Trans. A.S.M.E., Vol. 30, 

p. 801). 

TABLE 24. 



POUNDS OF AIR PER POUND OF OIL AND RATIO OF AIR SUPPLIED TO THAT 






CHEMICALLY REQUIRED. 








Light Oil, 


Medium Oil, 


Heavy Oil, 




C, 84%; H 


, 13%; S, 0.8%; 


C, 85%; H 


, 12%; s, 0.8%; 


C, 86%; H 


, n%; S, 0.8%; 


Per Cent CO2 


N, 0.2%; 


. 1%; H2O. 1%. 


N, 0.2%; 


1%; HjO, 1%. 


N, 0.2%; 


, 1%; H2O, 1%. 


by Volume as 














Shown by 














Analysis of Dry 


Lb. of 

Air per 

Lb. of Oil. 


Ratio of Air 


Lb. of 

Air per 

Lb. of Oil. 


Ratio of Air 


Lb. of 

Air per 

Lb. of Oil. 


Ratio of Air 


Chimney Gases. 


Supply to 

Chemical 

Requirements. 


Supply to 

Chemical 

Requirements. 


Supply to 

Chemical 

Requirements. 


4 


51.40 


3.607 


51.93 


3.704 


52.45 


3.803 


5 


41.31 


2.899 


41.71 


2.975 


42.12 


3.054 


6 


34.58 


2.427 


34.90 


2.490 


35.23 


2.554 


7 


29.77 


2.089 


30.04 


2.143 


30.31 


2.198 


8 


26.17 


1.836 


26.39 


1.883 


26.62 


1.930 


9 


23.37 


1.640 


23.56 


1.680 


23.75 


1.722 


10 


21.12 


1.482 


21.29 


1.518 


21.45 


1.555 


11 


19.83 


1.391 


19.43 


1.386 


19.58 


1.419 


12 


17.76 


1.246 


17.88 


1.276 


18.01 


1.306 


13 


16.46 


1.155 


16.57 


1.182 


16.69 


1.210 


14 


15.36 


1.078 


15.45 


1.102 


15.55 


1.127 


15 


14.39 


1.010 


14.48 


1.033 


14.57 


1.056 



TABLE 25. 

BOILER EFFICIENCY FOR EXCESS AIR SUPPLY (OIL FUEL). 



Excess Air Supply, Per Cent. 



Assumed temperature of escaping gases, 
deg. fahr 

Corresponding ideal efficiency of boiler, 
per cent 

Possible saving in fuel due to reduction 
of air supply to 10 per cent excess, ex- 
pressed as per cent of oil actually 
burned under assumed conditions .... 



10 


50 


75 


100 


150 










Over 


400 


450 


475 


490 


500 
Under 


84.2 


80.27 


77.66 


75.22 


70.94 
Over 





4.67 


7.78 


10.68 


15.76 



Over 
500 
Under 

67.09 



Over 
20.32 



Table 26 gives the results of a series of tests made at the Redondo 
plant of the Pacific Light & Power Company, CaUfornia, on a 604-horse- 
power B. & W. boiler equipped with Hammel furnaces and burners. 
The boiler was in regular service and under usual operating conditions. 



FUELS AND COMBUSTION 



93 









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a 33 






94 STEAM POWER PLANT ENGINEERING 

50. Comparative Evaporative Economy of Oil and Coal. — In deter- 
mining the comparative economy of coal and oil, the fixed and operating 
charges must be considered in addition to the cost and efficiency of 
the fuel. From the market quotation on oil and coal and the com- 
parative heating values of each the actual cost per B.t.u. is readily 
obtained, and by combining this with the relative efficiencies from the 
furnace standpoint the net cost of the fuel is obtained. The fixed 
charges vary with the location and size of the plant and are approxi- 
mately the same per boiler horsepower for a given location in both 
cases. The insurance rates may be greater with the oil fuel and the 
depreciation of the boiler setting may be somewhat larger, but in a well- 
constructed furnace the latter item should be the same in both instances 
for average rates of combustion. The operating charges are decidedly 
in favor of the oil fuel, since no ash handling is necessary. Oil fuel is 
readily fed to the furnace, and the cost of attendance may be materially 
less than with coal firing, and one man may safely control from eight to 
ten boilers. 

51. Oil Burners. — The function of the burner is to atomize the oil 
to as nearly a gaseous state as possible. 

Classification of a few well-known burners 

Mechanical Spray: Spray Burners: 

Korting. Outside Mixers. 

a. Peabody. 
Vapor or Carburetor: 6. Warren. 

Durr. Inside Mixers. 

Harvey. a. Hammel. 

h. Kirkwood. 

c. Branch. 

d. Williams. 

Oil burners for burning liquid fuel may be divided into three general 
classes : 

1. Mechanical spray, in which the oil, previously heated to a temper- 
ature of about 150 deg. fahr., is forced under pressure through nozzles 

so designed as to break it up 
into a fine spray. The Korting 
Liquid Fuel Burner, Fig. 20, is 
an example of this type. In this 
design a central spindle, spirally 
Fig. 20. Korting Fuel Oil Burner. g.^Q^ed, imparts a rotary motion 

to the oil and causes it to fly into a spray by centrifugal force on 
issuing from the nozzle. The particles of oil are burned in the furnace 




FUELS AND COMBUSTION 95 

when they come in contact with the necessary air to effect combustion. 
This type of burner is httle used in this country in connection with 
power-plant work, but is meeting with much success in Europe. 

2. Vapor burners, or carburetors, in which the oil is volatilized in 
a heater or chamber and then admitted to the furnace, are seldom used 
except in connection with refined oils, as the residuals from crude oil are 
vaporized only at a high temperature. The Durr and Harvey gasifiers 
are the best known of this type. 

3. Spray burners are by far the most common in use. In this type 
the oil is held in suspension and forced into the furnace by means of a 
jet of steam or compressed air. Spray burners are designed either as 
outside mixers, in which the oil and atomizing medium meet outside 
the apparatus, or inside mixers, in which the oil and atomizing medium 
mingle inside the apparatus. 

The Peahody burner. Fig. 21, illustrates the principles of the "outside- 
mixer" type of apparatus. In this type the oil flows through a thin 
slit and falls upon a jet of steam which atomizes it and forces it into 
the furnace in a fan-shaped spray. A feature of this apparatus is its 
simplicity of construction. 

Fig. 22 illustrates the Hammel burner as used at the power house of 
the Pacific Light and Power Company, Los Angeles, Cal. Oil enters the 
burner under pressure and flows through opening D to the mouth of the 
burner, where it is atomized by the steam jets issuing from slots G, H, 
and 7. The oil is preheated to facilitate its flow through the supply 
system. Plates K-K are removable and are easily replaced when worn 
out or burned. The Hammel burner belongs to the "inside mixers." 

A few well-known types of "inside mixers" are illustrated in Figs. 22 
to 24. The operation is practically the same in all of them and they 
differ only in mechanical details. 

The simplest and most reliable burners are of the Hammel type and 
are much in evidence in the Pacific States. 

53. Furnaces for Burning Fuel Oil. — The efficient combustion of 
oil fuel depends more upon the proportions of the furnace than upon 
the type of burner, provided, of course, the latter is of modern design. 
While it is desirable to have incandescent brickwork around the flame 
it is impossible to do so in many cases and a satisfactory compromise 
is effected by using a flat flame burning close to a white-hot floor through 
which air is steadily flowing. A good burner will maintain a suspended 
flame clear and smokeless in a cold furnace. The path of the flame 
in the furnace must be such as to insure uniform distribution of heat 
over the boiler heat-absorbing surfaces without direct flame impinge- 
ment. Under ordinary firing the flame should not extend into the 



96 



STEAM POWER PLANT ENGINEERING 




jm 



Fig. 21, Peabody Fuel-oil Burner. 




qAQ/h 



Fig. 22. Hammel Fuel-oil Burner. 




Steam 



idlQT 



1 



Fig. 23. Branch Fuel-oil Burner. 



FUELS AND COMBUSTION 



97 




Fig. 24. Kirkwood Fuel-oil Burner. 




Fig. 25. Billow Type of Fuel-oil Burner. 





i^ 



M Oil Pipe 



■J4 Steam Pipe 





Fig. 26. Warren Fuel-oil Burner. 



98 



STEAM POWER PLANT ENGINEERING 



tubes. The first pass of the boiler should be located directly over the 
furnace in order that the heating surface may absorb the radiant energy 
from the incandescent fire brick. Fire-brick arches and target walls 
are not to be recommended on account of the localization of heat re- 
sulting in burning out the tubes or bagging the shell and on account of 
the hmited overload capacity. 




Fig. 27. Furnace for Burning Fuel Oil, Rear Feed (Hammel). 

Fig. 27 shows the general details of a Hammel oil-burning furnace 
illustrating current practice on the Pacific coast. The burner tip is 
housed in a slot located in the back of an arched recess in the bridge 
wall and the flame is projected forward toward the front of the furnace. 
The furnace floor is carried on pieces of old two-inch pipe or on old 



FUELS AND COMBUSTION 



99 



rails and is solid except for narrow air slots through the deck and in 
front of each arch. Each burner with its accompanying recess has a 
separate air tunnel from the boiler front; these tunnels do not commu- 
nicate with each other under the furnace floor and by closing the ash- 
pit door any tunnel can be sealed up while the others arc supplying air 
to their particular burners. The Hammel furnace is a modification of 
the well-known Peabody furnace, a section through which is shown in 
Fig. 28. 




Ficj. 28. Peabody Fuel-oil Furnace. 




Fig. 29. Modern Furnace for Burning Fuel Oil, Front Feed. 



100 



STEAM POWER PLANT ENGINEERING 



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FUELS AND COMBUSTION 101 

Fig. 29 gives the general details of a modern oil-burning furnace, 
with front feed, as applied to a horizontal return tubular boiler. 

53. Atomization of Oil. — For efficient combustion the oil should 
be injected into the furnace in the form of a spray. Three systems of 
atomization are in use in stationary practice, namely, mechanical, air, 
and steam. Of these, by far the greater number of installations in the 
United States are of the last order. 

The mechanical or Korting system is not much in evidence in this 
country, but it is used extensively in Europe. The operation of this 
system is described in paragraph 51. The makers state that to oper- 
ate the pumps and supply the heat to the oil takes from | to 1 per cent 
of the steam evaporated. Mechanical atomization presents many pos- 
sibihties and it is not unlikely that future development may lie along 
this path. 

In air atomization the air is used at pressures from 1| to 60 pounds 
per square inch, depending upon the type of burner. From Table 27 
it will be seen that the total steam used to compress the air varies 
from 1.01 to 7.45 per cent of the total generated. For air atomization 
and with air pressures of from 20 to 30 pounds per square inch, J. H. 
Hoppes (Jour. A.S.M.E., Aug., 1911, p. 902) states that from 6 to 10 
cubic feet of air per minute per pound of oil burned will be required. 
Compressed air offers no opportunity for fuel saving over the use of 
steam direct in cases where steam is available. In certain industrial 
operations where high temperatures are essential the use of air is pre- 
ferred. When it is necessary to use high-pressure air the economy 
decreases with the increase in pressure, since the cost of each cubic foot 
of compressed air increases rapidly with the pressure, but its ability to 
atomize the oil does not increase proportionately. 

Steam is the most commonly used medium for atomizing the oil, 
since its use obviates complication and risk of interrupted service. 
The amount of steam required to atomize the oil varies from 0.15 to 
0.7 pounds per pound of oil, with an average of about 0.4 pounds. The 
steam consumption is generally stated in per cent of the total steam 
generated, but the results are misleading since the percentage factor 
depends largely upon the efficiency of the boiler. Table 27 gives the 
results of a number of tests of different types of burners with air and 
steam as atomizing mediums. 

54. Oil-feeding Systems. — Fig. 30 gives a diagrammatic arrange- 
ment of the piping commonly employed in feeding oil fuel to the burners. 
Steam-actuated oil pumps, installed in duplicate, draw the fuel from 
the supply tank and deliver it under pressure to the burners. The 
piping is cross-connected so that repairs can be made without inter- 



102 



STEAM POWER PLANT ENGINEERING 




FUELS AND COMBUSTION 103 

rupting the service. The oil is heated from the pump exhaust before 
it is suppUed to the burners. This should not be carried beyond the 
flash point of the oil used or there will be danger from carbon deposits 
in the supply pipe. A strainer is placed in the suction line between 
the storage tank and the oil-pressure pump to minimize clogging of 
the burner. In some instances strainers are also placed in the supply 
pipe between the heater and burner. The relief valve between the 
pumps and burners is set at a definite maximum oil pressure so as to 
prevent excessive pressure. The oil meter is for the purpose of check- 
ing the storage tank indicator. All oil piping is installed so that it 
can be drained back to the storage tank by gravity in case of neces- 
sity. In many large plants the strainers, meters, heaters and piping 
are installed in duplicate. Arrangements are usually made for the oil 
to be delivered at constant pressure. The supply of steam to the 
burner is controlled by regulating the pressure in a separate main 
common to all burners, the pressure in the main bearing a certain pre- 
determined relation to the pressure in the oil mains. In most installa- 
tions the supply of steam and oil at the burner is regulated by hand 
to meet the requirements of the individual burners. At the Redondo 
plant of the Pacific Light and Power Company, Redondo, Cal., the 
supply of oil and steam to all burners and the supply of air for com- 
bustion to any number of boilers are automatically controlled from a 
central point. For a description of this system see Trans. A.S.M.E., 
Vol. 30, p. 808. 

Low-pressure systems are ordinarily operated under standpipe 
pressures as in Fig. 31, which illustrates the arrangement of apparatus 
as advocated by the International Gas and Fuel Company. A steam 
pump B draws the oil from the buried tank through pipe Z and delivers 
it to the standpipe E. Thence it flows through pipe I to the burners 
under a head of about 10 feet. The pump runs constantly, the surplus 
oil flowing back to the tank through the pipe T. The oil is heated by 
the exhaust pipe Z\ The oil pump is provided with a device D having 
a piston connected by a chain with a cock S, which automatically opens 
when the boiler is not under steam pressure, so that the standpipe will 
be emptied, the oil flowing to the storage tank. 

Fig. 32 illustrates the Hydraulic Oil Storage Company's system of 
storing oil and delivering it to the burners. The oil reservoirs are 
placed below grade, as indicated, to minimize fire risk. The operation 
is as follows: Water enters the ''float box" and flows through a "three- 
way cock" to the bottom of the reservoir until all of the oil and water- 
pipes are filled up to the level of the float box, when the float automati- 
cally cuts off the supply. This flooding of the entire system drives 



104 



STEAM POWER PLANT ENGINEERING 




Fig. 31. International Gas and Fuel Company's Fuel-oil System. 



SypbJOD 
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Inlei 




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Fig. 32. Hydraulic Oil Storage Company 's Fuel-oil System. 



I 



FUELS AND COMBUSTION 105 

out all of the air. The three-way cock is then turned to ''discharge" 
and part of the water flows to the sewer. The tank car or wagon is 
next attached to the "oil inlet" and the oil flows into the tank and dis- 
places the water until the level of the "filler float" is reached, when the 
supply is automatically cui off. The inlet is so placed that the head 
of oil in the tank car is sufficiently great to overcome the opposing 
head of water. The three-way valve is next turned to the first position 
and the head of water forces the oil to the burners. After the oil has 
been withdrawn from the storage tank the water can only rise to the 
level of the water in the float box and therefore cannot be fed to the 
furnace. The small steam pipe admits steam into the tank and heats 
the oil, thereby making it flow more freely. 

55. Oil Transportation and Storage. — Fuel oil is delivered in bulk, 
either in tank cars, barges or steamships, or by pipe lines, depending 
upon the location of the plant. It must be stored in accordance with 
underwriters' requirements and community ordinances. The require- 
ments of the National Board of Fire Underwriters as regards the storage 
and use of fuel oil are substantially as follows: 

All oil used for fuel purposes under these rules shall show a flash test 
of not less than 150 deg. fahr. (Abel-Pensky flash-point tester). This 
flash point corresponds closely to 160 deg. fahr. (Tagliabue open-cup 
tester), which may be used for rough estimations of the flash point. 

In closely built-up districts or within fire limits, tanks to be located 
underground with their tops not less than 3 ft. below the surface of 
the ground and below the level of the lowest pipe in the building to be 
supphed. Tanks may be permitted underneath a building if buried 
at least 3 ft. below the basement floor, which is to be of concrete not 
less than 6 in. thick. Tanks shall be set on a firm foundation and 
surrounded with soft earth or sand, well tamped into place. No air 
space shall be allowed immediately outside of tanks. The tanks may 
have a test well, provided the test well extends to near the bottom of 
the tank, and the top end shall be hermetically sealed and locked ex- 
cept when necessarily open. When the tank is located underneath a 
building, the test well shall extend at least 12 ft. above the source of 
supply. The limit of storage permitted shall depend upon the location 
of tanks with respect to the building to be supplied and adjacent build- 
ings, the permissible aggregate capacity if lower than any floor, base- 
ment, cellar or pit in any building within the radius specified being as 
follows : 

Capacity. Radius. 

Unlimited 50 ft. 

20,000 gal 30 ft. 

6,000 gal 20 ft. 

1,500 gal 10 ft. 

*500 gal Less than 10 ft. 

* In this case the tank must be entirely incased in 6 in. of concrete. 



106 STEAM POWER PLANT ENGINEERING 

When located underneath a building no tank shall exceed a capacity 
of 9,000 gal., and basement floors must be provided with ample means 
of support independent of any tank or concrete casing. 

Outside of closely built-up districts or outside of fire limits, above- 
ground storage tanks may be permitted provided drainage away from 
burnable property in case of breakage of tanks is arranged for or suitable 
dikes are built around the tanks. 

When above-ground tanks are used, all piping must be arranged 
so that in case of breakage of piping the oil will not be drained from the 
tanks. This requirement prohibits the use of gravity feed from stor- 
age tanks. Above-ground tanks of less than 1,000 gal. capacity with- 
out dikes may be permitted in case suitable housings for the protection 
of the tanks against injury are provided. 

MATERIAL AND CONSTRUCTION OF TANKS 

Tanks must be constructed of iron or steel plate of a gauge depending 
upon the capacity as specified in the following: 

UNDERGROUND TANKS INSIDE SPECIFIED FIRE LIMITS. 

Or Within 10 Ft. of a Building When Outside Such Limits. 

/-.„„„ -x,, r.„i Minimum Thickness 

Capacity, Gal. ^j Material. 

1 to 560 14 U. S. gauge 

561 to 1,100 12 U. S. gauge 

1,101 to 4,000 7 U. S. gauge 

4,001 to 10,500 I U. S. gauge 

10,501 to 20,000 1^ U. S. gauge 

20,001 to 30,000 f U. S. gauge 

UNDERGROUND TANKS OUTSIDE SPECIFIED FIRE LIMITS. 

Provided the Tanks are 10 Ft. or More from a Building. 

Cinflritv Gal Minimum Thickness 

capacity, Ual. ^^ Material. 

1 to ,30 18 U. S. gauge 

31 to 350 16 U. S. gauge 

351 to 1,100 14 U. S. gauge 

1,101 to 4,000 7 U. S. gauge 

4,001 to 10,500 i U. S. gauge 

10,501 to 20,000 ^ U. S. gauge 

20,001 to 30,000 f U. S. gauge 

Tanks of greater capacity than 30,000 gal. must be made of propor- 
tionately heavier metal. All joints of tanks must be riveted and sol- 
dered, riveted and calked, welded or brazed together, or made by some 
equally satisfactory process. The shells of tanks must be properly 
reinforced where connections are made and all connections so far as 
practicable made through the upper side of tanks above the oil level. 
Tanks shall be thoroughly coated on the outside with tar, asphaltum 
or other suitable rust-resisting material. 

FILL AND VENT PIPES 

Each underground storage tank having a capacity of over 1,000 gal. 
must be provided with at least a 1-in. vent pipe extending from the top 
of the tank to a point outside the building, and to terminate at a point 



FUELS AND COMBUSTION 107 

at least 12 ft. above the level of the top of the highest tank car or other 
reservoir from which the storage tank may be filled. The terminal 
must be provided with a hood or gooseneck protected by a noncor- 
rodible screen and be placed remote from fire escapes and never nearer 
than 3 ft., measured horizontally and vertically, from any window or 
other opening. Vent pipes from two or more tanks may be connected 
to one upright, provided the connection is made at a point at least 1 ft. 
above the level of the source of supply. 

Tanks having a capacity of less than 1,000 gal. may be provided with 
combined fill and vent pipes so arranged that the fill pipe cannot be 
opened without opening the vent pipe, these pipes to terminate in a 
metal box or casting provided with a lock. Fill pipes for tanks which 
are installed with permanently open vent pipes must be provided with 
metal covers or boxes, which are to be kept locked except during filHng 
operations. Fill and vent pipes for tanks located under buildings are 
to be run underneath the concrete floor to the outside of the building. 

Suitable filters or strainers for the oil should be installed and pref- 
erably be located in the supply line before reaching the pump. Filters 
must be arranged so as to be readily accessible for cleaning. Feed 
pumps must be of approved design, secure against leaks and be ar- 
ranged so that dangerous pressures will not be obtained in any part of 
the sytem. It is further recommended that feed pumps be intercon- 
nected with pressure air supply to burners to prevent flooding. 

Glass gages, the breakage of which would allow the escape of oil, 
are to be avoided. If their use is necessary, they should have sub- 
stantial protection or be arranged so that oil will not escape if broken. 
Pet-cocks must not be used on oil-carrying parts of the system. 

Receivers or accumulators, if used, must be designed so as to secure 
a factor of safety of not less than 6 and must be subjected to a pressure 
test of not less than twice the working pressure. The capacity of the 
oil chamber must not exceed 10 gal. A pressure gage must be provided; 
also an automatic relief valve set to operate at a safe pressure and 
connected by an overflow pipe to the supply tank, and so arranged that 
the oil will automatically drain back to the supply tank immediately 
on closing down the pump. 

If standpipes are used, their capacity shall not exceed 10 gal. They 
must be of substantial construction, equipped with an overflow and so 
arranged that the oil will automatically drain back to the supply tank 
on shutting down the pump, leaving not over 1 gal., where necessary, 
for priming, etc. If vented, the opening should be at the top and may 
be connected with the outside vent pipe from the storage tank, above the 
level of the source of supply. 

Piping must be run as directly as possible and pitched toward the 
supply tanks without traps. Overflow and return pipes must be at 
least one size larger than the supply pipes, and no pipe should be less 
than J-in. pipe size. Connection to outside tanks should be laid below 
the frost line and not placed near nor in the same trench with other 
piping. 

Readily accessible shutoff valves should be provided in the supply 
line as near to the tank as practicable, and additional shutoffs installed 
in the main fine inside the building and at each oil-consuming device. 



108 STEAM POWER PLANT ENGINEERING 

Controlling valves in which oil under pressure is in contact with the 
stem shall be provided with a stuffing-box of hberal size, containing 
a removable cupped gland designed to compress the packing against 
the valve stem and arranged so as to facilitate removal. Packing 
affected by the oil must not be used. The use of approved automatic 
shutoffs for the oil supply in case of breakage of pipes or excessive leak- 
age in the building is recommended. 

56. The Purchase of Fuel Oil. — The following extracts from Bulletin 
No. 3, 1911, Bureau of Mines (''Specifications for the Purchase of Fuel 
Oil for the Government, with Directions for Sampling Oil and Natural 
Gas"), though primarily intended for the guidance of Government 
officials, may be of service to engineers: 

1. In determining the award of a contract, consideration will be 
given to the quality of the fuel offered by the bidders, as well as the 
price, and should it appear to be the best interest of the Government 
to award a contract at a higher price than that named in the lowest 
bid or bids received, the contract will be so awarded. 

2. Fuel oil should be either a natural homogeneous oil or a homo- 
geneous residue from a natural oil; if the latter, all constituents having 
a low flash point should have been removed by distillation; it should 
not be composed of a light oil and a heavy residue mixed in such pro- 
portions as to give the density desired. 

3. It should not have been distilled at a temperature high enough 
to burn it nor at a temperature so high that flecks of carbonaceous 
matter began to separate. 

4. It should not flash below 140 deg. fahr. in a closed Abel-Pensky 
or Pensky-Martens tester. 

5. Its specific gravity should range from 0.85 to 0.96 at (59 deg. 
fahr.); the oil should be rejected if its specific gravity is above 0.97 
at that temperature. 

6. It should be mobile, free from solid or semi-solid bodies, and should 
flow readily at ordinary atmospheric temperatures and under a head of 
1 foot of oil, through a 4-inch pipe 10 feet in length. 

7. It should not congeal or become too sluggish to flow at 32 deg. 
fahr. 

8. It should have a calorific value of not less than 10,000 calories 
per gram (18,000 B.t.u. per pound); 10,250 calories to be the standard. 
A bonus is to be paid or a penalty deducted according to the method 
stated under section 21, as the fuel oil delivered is above or below this 
standard. 

9. It should be rejected if it contains more than 2 per cent water. 

10. It should be rejected if it contains more than 1 per cent sulphur. 

11. It should not contain more than a trace of sand, clay or dirt. 

12. Each bidder must submit an accurate statement regarding the 
fuel oil he proposes to furnish. This statement should show: 

a. The commercial name of the oil. 

b. The name or designation of the field from which the oil is obtained. 



FUELS AND COMBUSTION 109 

c. Whether the oil is a crude oil, ti refinery residue, or a distillate. 

d. The name and location of the refinery, if the oil has been refined 
at all. 

For sampHng, analysis, etc., consult complete bulletin. 

Analyses of California Petroleums: Bulletin No. 19, U. S. Bureau of Mines, 1912. 

Atomization: Jour. A.S.M.E., Aug. 11, 1911, p. 883, 902; Jour. El. Power and Gas, 
Dec. 23, 1911. 

Burners: Jour. El. Power and Gas, Dec. 23, 1911, Apr. 1, 1911; Engng., Feb. 16, 
1912; Power, Jan. 27, 1914, p. 139. 

Comparative Evaporative Value of Coal and Oil: Jour. El. Power and Gas, March 
18, 1911; Jour. A.S.M.E., Aug. 11, 1911, p. 872. 

Draft Requirements for Burning Oil Fuel: Jour. A.S.M.E., Aug., 1911; Oct., 1912. 

Economy Tests with Oil Fuels: Trans. A.S.M.E., 30-1908, p. 775; Jour. A.S.M.E., 
Aug. 11, 1911, p. 940. 

Furnaces for burning Oil Fuel: Jour. El. Power and Gas, Dec. 30, 1911, Apr. 8, 
1911; Jour. A.S.M.E., Aug., 1911, p. 879; Ir. Td. Review, June 3, 1908; Power, 
June 16, 1908. 

Oil Fuel: Prac. Engr., July 15, 1916, p. 607. 

Oil for Steam Boilers: Jour. A. S.M.E,, Aug., 1911, p. 931; Jour. El. Power and Gas, 
Dec. 16, 1911; Power, Aug., 1908, p. 943; Jan. 23, 1908, p. 980; Bulletin No. 
131, Louisiana State University. 

Precautions with Oil Fuel: Eng. and Min. Jour., Apr. 1, 1911, p. 653. 

Purchase of Fuel Oil for the Government: Bulletin No. 3, Bureau of Mines, 1911. 

Regulation of Oil Supply to Burners: Trans. A.S.M.E., 30-1908, p. 804. 

Storage and Transportation: Jour. El. Power and Gas, Dec. 16, 1911, p. 564; Eng. 
News, Sept. 25, 1902, p. 232; Power, July 16, 1908. 

Unnecessary Losses in Firing Fuel Oil: Trans. A.S.M.E., 30-1908, p. 797. 

57. Gaseous Fuels. — The most commonly used gaseous fuels for 
steam generating purposes are natural gas, blast furnace gas and by- 
product coke oven gas. 

Natural gas is an ideal fuel for steam generation and offers all of the 
advantages of solid and Hquid fuels and none of the disadvantages. 
No storage bins or reservoirs are necessary, ashes are absent and standby 
losses may be reduced to a minimum. In the immediate locahty of 
natural gas wells, gas fired furnaces may prove to be more economical 
than coal furnaces but the limited supply restricts its use as a general 
fuel. A large combustion space is essential and a volume of 0.75 cubic 
feet per rated boiler horsepower will be found to give good results. The 
best results are obtained by employing a large number of small burners, 
each capable of handling 30 nominal rated horsepower. The use of 
a number of small burners obviates the danger of stratification of the 
gases which might occur with the large burners. A typical burner is 
illustrated in Fig. 33. A satisfactory working pressure is about 8 
ounces at the entrance of the burner. Table 28 gives typical analyses 
and calorific values of natural gas. 



no 



STEAM POWER PLANT ENGINEERING 



Although hlast-furnace gas is used extensively in gas engines its 
application to steam power generation is by no means discontinued, 
since the first high cost of a gas engine equipment, space requirements 
and high maintenance and attendance charges may more than offset 



I Gas Suppljr 



Gi For lighTloads 
GilGa Foe normal loads 
Gi,G3 For.heavy loads 
Gi,G2.G3 For heavy overloads 
K For low draft, low gas pressure 
or very heavy^ overloads 



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Detachable Nozzle 



Mixing Chamber LJ< Auxiliary Steam or 

Compressed Air 

Fig. 33. Gwynne Improved Gas Burner. 

the high thermal efficiency. A furnace volume of approximately 1 
to 1.5 cubic feet per rated boiler horsepower gives satisfactory results. 
The burner illustrated in Fig. 33 may be adapted to blast furnace gas 
by increasing the size of gas openings. On account of the possibifity 
of a pulsating action of the gases and the resulting puffs or explosions, 
settings for this class of work should be carefully constructed and thor- 
oughly buck staved. Blast furnace gas is very dirty and ample pro- 
vision should be made for removing the dust not only from the furnace 
but from the setting as a whole. Table 28 gives the chemical constit- 
uents and physical characteristics of a typical blast furnace gas. 

By-product coke oven gas has a higher calorific value than blast fur- 
nace gas but requires about the same type of burner and furnace design 
as the latter. It is ordinarily burned under a pressure of four inches 
of water. By-product coke oven gas is saturated with water vapor 
as it leaves the oven and provision should be made for removing the 
water of condensation before it reaches the burner. Tar and other 
hydrocarbons, which are present in considerable amount, tend to de- 
posit in the burners and render them inoperative. Accumulation of 
this deposit is ordinarily prevented by ''blowing out" the burners with 
steam. 

The Utilization of Waste Heat for Steam Generating Purposes, Jour. A.S.M.E., 
Nov., 1916, p. 859. 

Fig. 34 shows a section through a small experimental boiler designed 
by Prof. Wm. A. Bone, University of Leeds, England, which involves 
the principle of so-called "surface combustion," and for which ex- 
travagant claims have been made as regards efficiency and capacity. 



FUELS AND COMBUSTION 



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112 



STEAM POWER PLANT ENGINEERING 



It consists essentially of a plain tubular boiler, having ten tubes, 3 
inches in internal diameter. Each of these is bushed with a short tube, 
E, of fire clay and is filled for the rest of its length with finely broken 
refractory material. Mixing chambers of special design are attached 




Fig. 34. Experimsntal Boiler Involving the Principles of "Surface Combustion." 

to the front plate of the boiler as indicated. The mixture fed into the 
boiler tubes from these mixing chambers consists of the combustible 
gas with a proportion of air very slightly in excess of that theoretically 
required for combustion. The mixture is injected or drawn in through 
the orifice in the fire-clay plug. The gas burns without flame in the 
front end of the tube, the incandescent mass being in direct contact 
with the heating surface. The combustion of the mixture in contact 
with the incandescent material is completed before it has traversed a 
length of 6 inches from the point of entry of the tube. Although the 
core of the material at this part of the tube is incandescent the heat 
transference is so rapid that the walls of the tubes are considerably 
below red heat. The evaporation in regular working order is over 
20 pounds per square foot of heating surface and this can be increased 
50 per cent with a reduction in efficiency of only 5 or 6 per cent. The 
feures given by Prof. Bone for the boiler and economizer are as follows : 

Date, Dec. 8, 1910. 

Pressure of mixture entering boiler tubes, inches of water 17.3 

Pressure of products entering economizer, inches of water 2.0 

Steam pressure, pounds per square inch gauge 100 . 

Temperature of steam in boiler, deg. fahr 334.0 

Temperature of gases leaving boiler, deg. fahr 446.0 

Temperature of gases leaving economizer, deg. fahr. 203 . 

Temperature of water entering economizer, deg. fahr 41.9 



FUELS AND COMBUSTION 113 

Temperature of water leaving economizer, deg. fahr 13G.4 

Evaporation per square foot of heating surface per hour, pounds. . 21 .6 
Gas consumption, cubic feet per hour, at 32 deg. fahr. and 14.7 

pounds per square inch 996 . 

B.t.u. per standard cubic foot (lower heat value) 562.0 

Water evaporated per hour from and at 212 deg. fahr., pounds. . . . 550.0 
Efficiency of boOer and economizer (on basis of low heat value), per 

cent 94.3 

For further details of Prof. Bone's experiment see American Gas Light Journal, 
Dec. 4, 1911; Engineering, April 14, 1911; Engineer (London), April 14, 1911. 
See also editorial, Industrial Engineering, Jan., 1912, p. 59. 

At this writing (1917) practically nothing has been done in this country toward 
applying Prof. Bone's principles to steam boilers. For results of recent work in 
surface combustion see Power, Feb. 13, 1917, p. 225. 

Burning Natural Gas under Boilers: Power, Oct. 22, 1912, p. 897. 
Coke Oven Gas as a Fuel: Power, Aug. 29, 1916, p. 310. 



PROBLEMS. 

1. The following analyses were obtained from a sample of Illinois coal "as re- 
ceived ' ' : 

Proximate Analysis. Ultimate Analysis. 

Per Cent. Per Cent, 

Moisture 12.39 Hydrogen 5.85 

Volatile matter 36.89 Carbon 61.29 

Fixed carbon 41 . 80 Nitrogen 1 . 00 

Ash 8.92 Oxygen 19.02 

100.00 Sulphur 3.92 

Ash 8.92 

100.00 

a. Transfer these analyses to the "moisture free" and "moisture and ash free" 
basis. 

h. Transfer the ultimate analysis to the "moisture, ash and sulphur free" basis. 

c. Determine the free "hydrogen," "combined moisture" and "total mois- 
ture." 

d. Calculate the ultimate analysis from the proximate analysis. 

2. If the moisture and ash contents of an Illinois coal are 8 per cent and 12 per 
cent respectively, approximate the ultimate analysis by Evans' method. (See 
Example 4.) 

3. Using Dulong's formula calculate the calorific value of the dry coal as per 
analysis given in Problem 1. 

4. Using Dulong's formula approximate the calorific value of the coal as received 
considering the calculated values of the ultimate analysis. (See Example 5.) 

5. Using the data in Problem 1, calculate the theoretical air requirements per 
pound of coal as fired. 

6. Required the character and amount (by weight) of the products of combustion 
resulting from the complete combustion of the coal designated in Problem 1 with 
theoretical air requirements. 



114 STEAM POWER PLANT ENGINEERING 

7. Same data as in Problem 6. Determine the per cent by volume of the CO2 
in the flue gas. 

8. Determine the weight of dry air supplied per pound of coal as fired, analysis 
as in Problem 1, if the flue gas resulting from the combustion is composed of 

CO2 13.00 O2 5.30 

CO 0.44 No 81.26 

(Per cent by volume) 

9. Calculate the theoretical temperature of combustion if the coal as fired, analy- 
sis as in Problem 1, is completely burned with 50 per cent air excess. 

10. If coke breeze containing 85 per cent carbon and 15 per cent ash is completely 
burned under a boiler with 50 per cent air excess and the flue gas temperature is 
500 deg. fahr., required the heat loss in the flue gas per lb. of fuel as fired if the 
temperature of the air supply is 80 deg. fahr. 

11. If the flue gas resulting from the combustion of the fuel designated in Problem 
10 contains 0.5 per cent CO and 12 per cent CO2 (by volume), required the loss due 
to incomplete combustion of the carbon. 

12. Calculate the heat loss in the refuse if the coal as fired has an ash content of 
15 per cent and the combustible in the dry refuse is 20 per cent of the dry refuse. 
Calorific value of the combustible in the ash, 13,600 B.t.u. per lb. 

13. Required the heat lost per lb. of coal as fired in evaporating the moisture 
from the coal designated in Problem 1 if the temperature of the flue gas is 500 deg. 
fahr. and that of the boiler room, 80 deg. fahr. 

14. If crude oil containing 14 per cent of hydrogen and 3 per cent of oxygen is 
burned under a boiler, required the amount of heat lost per lb. of oil due to the for- 
mation of water by the combustion of the hydrogen. Flue gas temperature, 450 
deg. fahr., temperature of the oil, 120 deg. fahr. 

15. The following data were obtained from a boiler evaporation test: 

Heat absorbed by the boiler, 70 per cent of the calorific value of the coal as fired. 
Analysis of the coal as fired: 

Per Cent. Per Cent. 

Carbon 65 Ash and Sulphur 12 

Oxygen 8 Free moisture 10 

Hydrogen 4 Nitrogen . . . .' 1 

Calorific value as fired, 10,350 B.t.u. per lb. 
Flue Gas Analysis: 

Per Cent. Per Cent. 

CO2 14.18 CO 1.42 

O2 3 . 55 N2 80. 85 (by difference) 

Temperature of air entering furnace, 80 deg. fahr., temperature of the flue gas, 
480 deg. fahr., temperature of the steam in the boiler, 350 deg. fahr., relative humidity 
of the air entering the furnace, 70 per cent, combustible in the dry refuse, 20 per cent. 
a. Calculate the actual losses in per cent of the coal as fired. 
6. Calculate the inherent losses in per cent of the coal as fired, 
c. Approximate the extent to which the actual losses may be reduced by careful 
operation and proper design. 

16. 800,000 pounds of water are fed into a 72-inch by 20-ft. return tubular boiler 
during a period of 30 days; total weight of coal fed to furnace, 150,000 lb; coal used 
in banking fires and in starting up, 35,000 lb.; water "blown off," 1 gauge (4-in.) 
per day; boiler pressure, 115 lb. per sq. in. gauge; required the extent of the stand- 
by losses in per cent of the net actual evaporation. 



CHAPTER III 

BOILERS 

58. General. — The modern steam boiler is substantially identical 
in general design to its prototype of a generation ago. Increased 
pressure and superheat have necessitated improvements in structural 
details but changes affecting safety and operation are not to be confused 
with changes of design. Many of the small hand-fired boilers of today 
are in every way identical with those of twenty years ago. Great 
improvements have been made in the design and construction of the 
furnace and setting, mechanical stokers have been perfected and the 
maximum size of units has been vastly increased but the boiler proper 
is basically unchanged. 

As affecting fuel economy the boiler equipment is by far the most 
important part of the power plant and involves the largest share of the 
operating expenses. It matters little how elaborate, modern, or well 
designed it may be, skill, good judgment, and continued vigilance are 
required on the part of the operator to secure the best efficiency. 

Of the various types and grades of boilers on the market experience 
shows that most of them are capal)le of practically the same evapo- 
ration per pound of coal, provided they are designed with the same 
portions of heating and grate surface and are operated under similar 
conditions. They differ, however, with respect to space occupied, 
weight, capacity, first cost, and adaptability to particular conditions 
of operation and location. 

There is a tendency towards standardization of boiler construction 
by legislation in various communities. In view of the numerous adop- 
tions of the ''Standard Specifications for the Construction of Steam 
Boilers and Other Pressure Vessels"* as formulated by the American 
Society of Mechanical Engineers, it is not unlikely but that this code 
will ultimately be the required standard for all communities in the United 
States, t 

59. Classiflcation. — As to design and construction there is an almost 
endless variety of boilers and furnaces, classified as internally and 

* Trans. A.S.M.E., Vol. 36, 1914, p. 981. Also printed in pamphlet form. 

t These rules do not apply to boilers which are subject to federal inspection and 
control, including marine boilers, boilers of steam locomotive and other self-])ro- 
pelled railroad apparatus. 

115 



116 



STEAM POWER PLANT ENGINEERING 



externally fired; water tube snidfire tube; through tube and return tubular; 
horizontal and vertical. 

The internally fired type includes the vertical tubular, locomotive, 
Scotch-marine, and practically all flue boilers. The externally fired 
includes the plain cylinder, the through tubular, return tubular, and 
nearly all stationary water-tube boilers. A few well-known types will 
be described in detail. 

60. Vertical Tubular Boilers. — Figs. 1 and 35 illustrate typical 
fire-tube boilers of the internally fired class. The type shown in Figs. 
1 and 35 are commonly used where small power, compactness, low first 

cost and sometimes portability are 
chief requirements. They are seldom 
constructed in sizes over 100 horse- 
power. The tubes are placed sym- 
metrically with a continuous clear 
space between them and the spaces 
crossing the tube section at right an- 
gles, by means of which the tubes and 
tube sheet may be readily cleaned. 
Two styles are in common use; the 
exposed tube, Fig. 1, and the submerged 
tube. In the former the tube sheet 
and the upper portion of the tubes 
are exposed to the steam and in the 
latter they are completely submerged. 
According to the A.S.M.E. Boiler 
Code, not less than seven hand holes 
or wash out plugs are required for 
boilers of the exposed tube type; 
three in the shell at or about the line 
of the crown sheet, one in the shell at 
or about the fusible plug and three 
in the shell at the lower part of the 
water leg. In the submerged type 
two or more additional hand holes are required in the shell in line with 
the upper tube sheet. The distance between the crown sheet and the 
top of the grate should never be less than 24 inches even in the smallest 
boiler and should be as great as possible to insure good combustion. 

The advantages of this type of boiler are: (1) compactness and port- 
abihty; (2) requires no setting beyond a light foundation; (3) is a 
rapid steamer, and (4) is low in first cost. The disadvantages are: 
(1) inaccessibility for thorough inspection and cleaning; (2) small steam 




Fig. 35. Vertical Tubular Boiler 
with Submerged Tube Sheet. 



BOILERS 



117 



space, which results in excessive priming at heav>^ loads; (3) poor econ- 
omy except at light loads, as the products of combustion escape at a 
high temperature on account of the shortness of the tubes; (4) smoke- 
less combustion practically impossible with bituminous coals; (5) the 



small water capacity results in 



^ 



t 



i. ^ 




rapidly fluctuating steam pressures 
with varying demands for steam. 

Although vertical fire-tube boilers 
are usually of very small size, being 
seldom constructed in sizes over 100 
horsepower, an exception is found in 
the Manning boiler. Fig. 36, which is 
constructed in sizes as large as 250 
horsepower. Many of the disadvan- 
tages found in the smaller types are 
obviated in the Manning boilers, 
which, as far as safety and efficiency 
are concerned, rank with any of the 
other first-class types. They differ 
from the boiler described above 
mainly in having the lower or fur- 
nace portion of much greater diameter 
than the upper part which encircles 
the tubes. This permits a proper 
proportion of grate, which is not ob- 
tainable in boilers like Figs. 1 and 35. 
The double-flanged head connecting 
the upper and lower shells allows 



Fig. 36. Manning Vertical Fire-tube Boiler 



118 



STEAM POWER PLANT ENGINEERING 



sufficient flexibility between the top and bottom tube sheets to provide 
for unequal expansion of tubes and shell. The ash pit is built of brick 
and the water leg does not extend below the grate level, thus doing 
away with dead-water space. Where overhead room permits and 
ground space is expensive, this boiler offers the advantage of taking up 
a small floor space as compared with horizontal types. 

61. Fire-box Boilers. — Although vertical fire-tube boilers may be 
classed as fire-box boilers, yet the term ''fire box" is usually associated 
with the locomotive types, whether used for traction or stationary pur- 
poses. The usual form of fire-box boiler as applied to stationary work 
is illustrated in Fig. 37. The shell is prolonged beyond the front tube 



Safety Valve 




Fire DooF 



Asb Door 



Fig. 37. Typical Fire-box Boiler — Portable Type. 



sheet to form a smoke box. The front ends of the tubes lead into the 
smoke box and the rear ends into the furnace or fire box. The fire box 
is ordinarily of rectangular cross section, and is secured against collapse 
by stay bolts and other forms of stays. In Fig. 37 the smoke box is of 
cylindrical cross section and hence requires no staying except at the 
flat surface. Fire-box boilers are used a great deal in small heating 
plants where space limitation precludes other types. Their steam 
capacity gives them an advantage over the vertical tubular form. 
Being internally fired no brick setting is required. They are usually 
of cheap construction, designed for low pressure, and seldom made in 
sizes over 75 horsepower. Unless carefully designed and constructed 
high steam pressures are apt to cause leakage because of unequal ex- 
pansion of boiler shell, tubes, and fire box. Portable fire-box boilers 



BOILERS 



119 



with return tubes 
pressures as high 
than some of the 
where portabihty 
Fig. 38 shows b 
''brick set" type, 
pressure heating 
work and the 



are made in sizes as large as 150 horsepower and for 
as 150 pounds per square inch, but being more costly 
other types of boilers of equal capacity are used only 
is an essential requirement. 

t longitudinal section through a fire-box boiler of the 
. This class of boiler is much in evidence in lower 
installations. The boiler proper is encased in brick 
are carried along the outer surface of the shell 



Breeching Connection 



Repulator^ 
Water Line 



e 



^^ Safety Valve 



C— IStWi= 



-Steam Supply 

/-Layer of Martif 




Fig. 38. Typical Fire-box Boiler, " Kewanee Brickset. " 

before being discharged through the breeching. For smokeless com- 
bustion the furnace is fitted with a down-draft grate (see paragraph 98), 
and lined with refractory material. 

62. Scotch-marine Boiler. — Where an internally fired boiler is 
desired for large powers the Scotch-marine type is finding much favor 
with engineers. A number of the tall office buildings in Chicago are 
equipped with boilers of this class which are giving good results. They 
require little overhead room, no brick setting, and are excellent steamers. 
The Continental boiler, Fig. 39, is one of the best known of this type. 
The boiler is self-contained and requires no brick setting, the only fire 
brick used being those that form the bridge wall, baffle ring, and the 
layer at the back of the combustion chamber. The furnace and tubes 
are entirely surrounded by water, so that all fire surfaces, excepting the 
rear of the combustion chamber, are water cooled. The furnace is 
corrugated for its whole length. These corrugations, in addition to 
giving greater strength to the furnace, act as a series of expansion 
joints, taking up the strains due to unequal expansion of furnace and 
shell. Practically all types of mechanical stokers and grates are appli- 
cable to these boilers. The advantages of a Scotch boiler and of all 



120 



STEAM POWER PLANT ENGINEERING 



ponoag doj I ,^4 \f: ^,,r I 




BOILERS 



121 



internally fired boilers are: (1) minimum radiation losses; (2) require 
no setting; (3) no leakage of cool air into the furnace as sometimes 
occurs through cracks or porous brickwork of other types; (4) large 
steaming capacity for the space occupied. The circulation, however, 
is not always positive and the water below the furnace may be con- 
siderably below the average or normal temperature, giving rise to un- 
equal expansion and contraction which may cause leakage. The boiler 
proper is relatively costly, but this is offset to some extent by the 
absence of setting. 

63. Robb-Mumford Boiler. — Fig. 40 shows a section through a 
Robb-Mumford boiler, which is a modification of the Scotch-marine 
and of the horizontal tubular type. It consists of two cylindrical 




Fig. 40. Robb-Mumford Boiler. 



shells, the lower one containing a round furnace and tubes and the 
upper one forming the steam drum, the two being connected by two 
necks. The lower shell has an incline of about one inch per foot from 
the horizontal, for the purpose of promoting circulation and draft, 
and also for convenience in washing out the lower shell. Combustion 
takes place in the furnace, which is surrounded entirely by water, and 
the gases pass through the tubes and return between the lower and 
upper shells (this space being inclosed by a steel casing) to the outlet 
at the front of the boiler. Mingled water and steam circulate rapidly 
up the rear neck into the steam drum, where the steam is released, the 
water passing along the upper drum towards the front of the boiler and 
down the front neck, a semi-circular baffle plate around the furnace 
causing the down-flowing water to circulate to the lowest part of the 



122 STEAM POWER PLANT ENGINEERING 

lower shell under the furnace. The outer casing, which incloses the 
space between the lower and upper shells, including the rear smoke 
box and the smoke outlet, is constructed of steel plate, wdth angle-iron 
stiffeners, the various sections being bolted together for convenient 
removal. The inside of the steel case, including the rear smoke cham- 
ber, is Uned with asbestos air-cell blocks fitted in between the angle-iron 
stiffeners. The top of the upper drum and bottom of the lower shell 
are also covered with non-conducting material after the boiler is erected. 
Owing to the fact that steam and water spaces are divided between two 
cyhndrical shells, the thickness of plates is not so great as in the Scotch- 
marine or horizontal return tubular types; and the rear chamber of the 
marine boiler is avoided. 

The chief claim for this type of boiler is compactness. A battery of 
five 200-horsepower units occupies a floor space of but 33 feet in width 
by 20 feet in depth and 12.5 feet high. Each unit is entirely inde- 
pendent and may be isolated for cleaning, inspection, and repairs. 

64. Horizontal Return Tubular Boilers. — These are the most com- 
mon in use and are constructed in sizes up to 500 horsepower. They 
are simple and inexpensive and, when properly operated, durable and 
economical. Figs. 41 to 44 show various forms of standard settings, 
and Figs. 88, 89, and 90 different ''smokeless" settings. The grate is 
independent of the boiler, and the products of combustion pass beneath 
the shell to the back end, returning through the tubes to the front, 
and into the smoke connection. 

The tubes are from 3 to 4 inches in diameter and from 14 to 18 feet 
long, and are expanded into the tube sheets. The portion of the tube 
sheets not supported by the tubes is secured against bulging by suitable 
stays. Access to the interior of the boiler is obtained through manholes. 
The most convenient arrangement for inspection and cleaning is to have 
one manhole located at the top of the shell and one at the bottom of 
the front tube sheet. Return tubular boilers are made either with an 
extended or half-arch front (Fig. 41) or flush front (Fig. 42). The shell 
may be supported by lugs resting on the brickwork as in Fig. 41 or by 
steel beams and hangers as in Fig. 43. The latter construction permits 
the brickwork and shell to expand or contract independently, and settling 
of the brickwork does not affect the boiler alignment. According to the 
A.S.M.E. Boiler Code all horizontal tubular boilers over 78 in. in di- 
ameter are required to be supported by this outside suspension type of 
setting. With the side bracket support, the front lugs usually rest 
directly on iron or steel plates embedded in the brickwork, and the back 
lugs on rollers, to permit free expansion and contraction. The brackets 
are long enough to rest upon the outside wall, so that the inside brick 



BOILERS 



123 





124 



STEA]M POWER PLANT ENGINEERING 




I 



BOILERS 



125 



lining can be renewed without disturbing the setting. The A.S.M.E. 
Boiler Code specifies four pairs of brackets (two pairs on each side) 
for boilers over 54 in. and up to and including 78 in. in diameter, and 
not less than two brackets on each side for boilers up to and including 




Fig. 43. Return Tubular Boiler Setting — Outside Suspension Type. 



54 in. in diameter. The distance betw^een the rear tube sheet and wall 
should be about 16 inches for boilers less than 60 inches in diameter 
and from 20 to 24 inches for larger ones. The distance between grate 
and boiler shell should not be less than 28 inches for anthracite coal 
and 36 inches for bituminous coal.* The greater this distance the more 
complete the combustion, since the gases will have a better opportunity 
for combining with the air before coming into contact with the compara- 
tively cool surfaces of the shell. The shell should be slightly inclined 
toward the blow-off end so as to drain freely. 

The vertical distance between the bridge wall and shell is usually 
between 10 and 12 inches. The lower part of the combustion chamber 
behind the bridge wall may be filled with earth and paved with common 
brick as in Fig. 44 or left empty as in Fig. 42. The shape of the bridge 
walls whether curved to conform to the shell or flat appears to have 
little influence on the economy. 

The side and end walls are ordinarily constructed of common brick 
with an inner lining of fire brick, and may be sohd as in Fig. 42 or 
double with air spaces as in Fig. 41. The latter construction permits 
the inner and outer walls to expand independently without cracking 



* For smokeless combustion the setting must be modified, 
trated and described in paragraph 93. 



See furnaces illus- 



126 



STEAM POWER PLANT ENGINEERING 




BOILERS 



127 



and settling. Tests conducted by Ray and Kreisinger * show that a 
solid wall is a better heat insulator than a wall of the same total thickness 
containing an air space, hence if air spaces are used they should be filled 
with loose non-conducting material. The side walls are braced by 
five pairs of buckstaves, with through rods under the paving and 
over the tops of the boilers. Air leakage through the setting is com- 
pletely eliminated by enclosing the entire setting within a steel casing. 
A lining of kieselguhr or similar insulating material within the casing 
will greatly reduce the heat losses through the walls of the setting. 
See '' Insulation of Boiler Settings," Joseph Harrington, Power, Mar. 
27, 1917, p. 410. 

The connection between the rear wall and the shell is a source of 
more or less trouble on account of the expansion and contraction of the 
boiler. Cast-iron supports of T section supporting a fire-brick arch 




Fig. 45. Furnace Arch Bars 



Back Connection Made with 
Cast-iron Plate. 



are usually employed as illustrated in Fig. 45, the clearance between 
the arch and the shell being sufficient to allow the necessary expansion. 
In order to avoid air leakage this clearance space is filled with asbestos 
fiber. 

Fig. 46 shows the common method of resting one end of the arch 
supports on the rear wall and the other end on an angle iron riveted to 
the boiler. Fig. 47 illustrates the principles of the Woolson Arch 
connection. 

The products of combustion are sometimes carried over the top of 
the boiler as shown in Fig. 44. This tends to superheat the steam, but 
the advantage gained is probably offset considerably by the extra cost 
of the setting and the accumulation of soot on the top of the shell. The 
arrangement is not common. 

* Bui. No. 8, U. S. Bureau of Mines, 1911. 



128 



STEAM POWER PLANT ENGINEERING 



(-Brick on Edge 




Fig. 47. Woolson's Gas Tight Back 
Arch Connection. 



The steam connection is naturally made to the highest point in the 
boiler shell. Frequently a steam dome, to which the steam nozzle is 
connected, is provided as in Fig. 42. The function of the steam dome 
is to increase the steam space so as to permit the collection of dry 

steam at a point high above the 
water level. If a boiler is too small 
for its work and is forced far above 
its rating a steam dome is probably 
an advantage, though its use is less 
common now than formerly, since 
a properly designed boiler insures 
ample steam space without one. A 
dry pipe inside the boiler above the 
water hne as in Fig. 39 or 40 is 
commonly used to guard against 
priming where the nozzle is con- 
nected to the shell. 

For low pressures and smal-l powers 
the return tubular boiler has the ad- 
vantage of affording a large heating surface in a small space and large 
overload capacity. It requires little overhead room and its first cost 
is low. On the other hand the interior is difficult of access for purposes 
of cleaning and inspection. Boilers of this type are constructed in 
various sizes ranging from a 36-in. by 8 ft., rated at 15 horsepower, to a 
108-in. by 21 ft., rated at 500 horsepower, though sizes above 200 horse- 
power are exceptional. The working pressure seldom exceeds 150 
pounds per square inch. 

The standard externally fired return tubular boiler is limited in size 
since the damage from overheating the shell directly over the fire bed 
increases rapidly with the increase in thickness of the plate. The 
Lyons boiler overcomes this restriction through the addition of a bank 
of water tubes which form a roof to the furnace. These tubes protect 
the shell from the direct action of the gases and insure a positive and 
rapid circulation. They are covered with tile or spHt brick and form 
the equivalent of a ''Dutch oven." 

Arches— Firebrick Furnace: Jour. A.S.M.E., Jan., 1916, p. 7; Power, Feb. 20, 1912, 
Oct. 24, 1916, p. 598. 



65. Babcock & Wilcox Boilers. — Fig. 48 shows a longitudinal section 
through a Babcock & Wilcox boiler, illustrating a typical horizontal 
water-tube type. The tubes, usually 4 inciies in diameter and 18 feet 
in length, are arranged in vertical and horizontal rows and are expanded 



BOILERS 



129 



into pressed-steel headers. Two vertical rows are fitted to each header 
and are ''staggered" as shown in Fig. 49. The headers are connected 
with the steam drum by short tubes expanded into bored holes. Each 



steel Support 



Nozzl e 
Safety Valve 



Man-Hole 




rFloor 
Line 



^^^^^ Bridge Wall 

Fig. 48. Babcock & Wilcox Boiler and Standard Hand-fired Setting. 



tube is accessible for cleaning through openings closed by covers with 
ground joints held in place by forged steel clamps and bolts. The 
tubes are inchned at an angle of about 22 degrees with the horizontal. 
The rear headers are connected at the bottom to a 
forged steel mud drum. The steam drum is horizon- 
tal and the headers are arranged either vertically as 
shown in Fig. 94 or inchned as in Fig. 50. The 
boiler is supported by steel girders resting on suit- 
able columns independent of the brick setting. The 
grate is placed under the higher ends of the tubes, 
the products of combustion passing at right angles 
to the tubes and being deflected back and forth ])y 
fire-tile baffles. The feed water enters the front of 
the steam drum as shown in Fig. 50. A rapid cir- 
culation is effected by the difference in density be- 
tween the soHd column of water in the rear header 
and the mixed steam and water in the front one. 
Babcock & Wilcox boilers under 150 horsepower 
have but one steam drum, and the larger sizes have 




Fig. 49. Details of 
Header — Babcock 
& Wilcox Boiler. 



130 



STEAM POWER PLANT ENGINEERING 




Fig. 50. 



two. The drums are accessible for inspection through manhole open- 
ings. The number of tubes varies with the size of boiler, ranging from 
6 wide and 9 high in the 100 horsepower boiler to 14 high and 18 wide 
in the 500 horsepower boilers. The spacing of the tubes is based 

primarily upon the proportions of the 
grate. The width of the grate deter- 
mines the '' number of tubes wide" and 
the capacity of the boiler controls the 
''number of tubes high." Babcock & 
Wilcox boilers may be baffled so that 
the gases may pass out either at the 
front or rear of the top of the setting 
or at the rear of the bottom of the 
setting. The gases may be directed 
across the tubes as illustrated in Fig, 48 
or along the tubes as shown in Fig. 53.* 
Large doors in the sides of the setting 
give full access to all parts for inspec- 
tion and for removal of accumulations 
of soot. In the strictly modern power 
plant the setting is encased in steel in 
order to prevent air leakage, and the 
casing is lined on the inside with heat insulating material, such as 
kieselguhr, so as to reduce heat losses. 

Fig. 51 shows a section through a Babcock & Wilcox marine type 
boiler with cross drum. Boilers of this design have been installed 
in units of 1200 rated horsepower and are giving eminent satisfaction 
as to efficiency and capacity. For smokeless settings see Chapter IV. 
66. Heine BoUer. — Fig. 52 shows a longitudinal section through a 
Heine horizontal water-tube boiler. This boiler differs from the Bab- 
cock & Wilcox boiler in that the tubes are expanded into a single large 
header constructed of boiler steel. The drum and tubes are parallel 
with each other and inclined about 22 degrees with, the horizontal. The 
feed water enters at the front of the steam drum and flows into the mud 
drum, from which it passes to the rear header. Steam is taken from the 
front of the steam drum and is partially freed from moisture by the dry 
pipe A. A baffle over the front header prevents an excess of water 
from being carried into the dry pipe. As the rear header forms one 
large chamber, no additional mud drum is necessary and the sediment 
is " blown off " from the bottom by the blow-off cock. The circulation is 

* Horizontal and Vertical Baffling for B. & W. Boilers, S. H. Viall, Power, June 20, 
1916, p. 874. 



Front Section — Babcock 
& Wilcox Boiler. 



BOILERS 



131 




Fig. 51. Babcock & Wilcox Boiler — Cross Drum Type. 



Man-Hole 




Floor Line 



Bridge Wall 

Fig. 52. Heine Boiler and Standard Hand-fired Setting. 



132 STEAM POWER PLANT ENGINEERING 

somewhat freer than in the Babcock & Wilcox boiler on account of the 
large sectional area through the headers. 

Circulation in Water Tube Boilers: Jour. A.S.M.E., Jan. 1916, p. 17. 

67. Parker Boiler. — Fig. 53 shows a longitudinal sectional elevation 
and an end sectional elevation of a 1200-horsepower Parker down- 
flow boiler with double-ended setting. This type of boiler is finding 
much favor with engineers for central stations where large units are 
desired. The Parker boiler differs from the conventional horizontal 
water-tube boiler principally in circulation and flexibility. 

Feed water is pumped into the economizer or feed element (1), Fig. 
53, at 0, 0, and flows downward through a series of tubes, discharging 
finally into the drum through an upcast H. In a large unit, as illus- 
trated here, there are two feed elements and two drums. The circula- 
tion in the feed element is indicated by solid lines and arrow points at 
the left of the end sectional elevation, the tubes having been omitted 
from the drawing for the sake of clearness. 

The intermediate elements (2) take their water supply from the bot- 
tom of the drum through a cross-box V, the circulation being downward, 
as indicated by arrow points, through four tube wide elements, and 
finally discharge it through an upcast X into the steam space of the drum. 
Each element has a "down-comer" and an upcast. In the smaller- 
sized boilers the intermediate elements are omitted. 

The evaporator elements (3) take their water supply from the bottom 
of the drum at V, the circulation being downward through two tube 
wide elements, and finally discharge it into the drum at U. The last 
two passes of the water are through the two bottom tubes of each' 
element, thus assuring dry steam without the use of dry pipes. To 
prevent reversal of flow each element is fitted with a check valve at the 
admission end. Each drum is equipped with a diaphragm, as indicated, 
separating the steam and water spaces, thus insuring against foaming 
and priming. 

Saturated steam is taken from the drum at A and passes by way of 
B to C, where it enters the superheater S. The superheated steam 
leaves the superheater at D and passes by way of E and R and the 
storage drum N, finally leaving the boiler at G. The superheater is 
designed to maintain an approximately constant degree of superheat 
for all variations in load. 

All tubes are connected by malleable-iron junction boxes the interior 
of each tube being accessible through hand holes placed opposite the 
end of each tube. The hand-hole cover plates are on the inside of the 
box and have conical ground joints, thus dispensing with gaskets. 



BOILERS 



133 




134 



STEAM POWER PLANT ENGINEERING 



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SECTION THRU FURNACE 




FRONT ELEVATION 



LONGITUDINAL SECTION 



Fig. 54. Wickes Vertical Boiler with Steel Encased Setting. 



BOILERS 135 

The Parker boiler is built single or double ended, with or without 
superheater, and in sizes ranging from 50-horsepower to 2500-horse- 
power standard rating. 

68. Wickes Boiler. — Fig. 54 shows a section through a Wickes 
vertical boiler, illustrating the vertical water-tube type. The steam 
drum and water drum are arranged one directly above the other. The 
.tubes are expanded and rolled into both tube sheets and are divided 
into two sections by fire-brick tile. The water line in the steam drum 
is carried about two feet above the tube sheet, leaving a space of five 
feet between water line and top of the drum. This affords a large 
steam space and disengagement surface. Feed water is introduced 
into the steam drum below the water line and flows downward through 
the tubes of the second compartment. The boiler is supported by four 
brackets riveted to the shell of the bottom drum and is independent 
of the setting. The entire boiler is enclosed in a steel casing, insulated 
with non-conducting material and lined with fire brick. The boiler 
is completely surrounded by the products of combustion. The steel 
encased setting prevents lowering the temperature of the products 
of combustion by air infiltration and reduces radiation losses. The 
upper part of the steam drum acts as a superheating surface and tends 
to dry the steam. Wickes boilers are simple in design, easy to inspect 
and clean, low in first cost, and comparable in efficiency with any water- 
tube type of boiler. 

69. The Bigelow-Hornsby Boiler. — Fig. 55 shows a vertical section 
through a Bigelow-Hornsby boiler equipped with Foster superheater and 
Taylor stoker. This boiler is of the vertical water-tube type and is 
made up of a number of cylindrical elements, each element comprising 
an upper and lower drum connected by straight tubes. The two front 
elements are inclined over the furnace at an angle of about 68 degrees, 
and the two rear elements are vertical. The upper drums of the ele- 
ments are connected to a horizontal main steam drum by flexible tubing 
as indicated. Four elements constitute a section with an effective 
heating surface of 1250 square feet. Any number of sections may be 
connected together forming units of from 250 to 2500 boiler horsepower 
or more. All parts, both external and internal, are readily accessible. 
Feed water enters the top drum of the rear elements and passes twice 
the length of the tubes before entering into the general circulation. 
This arrangement permits a considerable portion of the impurities in 
the water to be precipitated in the rear drum from which they are 
readily discharged. By the time the water reaches the front of the 
boiler directly over the furnace, where the heat transmission is the most 
intense, the scale-forming elements have been practically eliminated. 



136 



STEAM POWER PLANT ENGINEERING 



The particular features of this boiler lie in the great extent of heating 
surface exposed to radiant heat and the height and volume of the com- 
bustion chamber. Bigelow boilers are productive of high economy 
and are readily forced to twice their rated capacity with little decrease 




Fig. 55. Bigelow-Hornsby Boiler and Setting. 



in over-all efficiency. The most notable installation of Bigelow boilers 
in this country is at the power plant of the Hartford Electric Light & 
Power Company, Hartford, Conn., where two 1250- and one 2500-boiler- 
horsepower units are installed. 

70. Stirling Boiler. — Fig. 56 shows a longitudinal section through a 
Stirling water-tube boiler, which differs considerably from the types 
just described. Three horizontal steam drums and one horizontal mud 
drum are connected by a series of inclined tubes. The tubes are bent 
at the ends to permit them to enter the drums radially. Short tubes 



f 



BOILERS 



137 



connect the steam spaces of all the upper drums and also the water 
spaces of the front and middle drums. Suitably disposed fire-tile 
baffles between the banks of tubes direct the gases in their proper 
course. The boiler is supported on a structural steel framework in- 




FiG. 56, Stirling Boiler and Standard Hand-fired Setting. 



dependent of the setting. The feed water enters the rear upper drum, 
which is the cooler part of the boiler, and flows to the bottom or mud 
drum, where it is heated to such an extent that many of the impurities 
are precipitated. There is a rapid circulation up the front bank of 
tubes to the front drum, across to the middle drum, and thence down 
the middle bank of tubes to the mud drum. The interior of the drums 
is accessible for cleaning by manholes located in the ends. The Stirling 
furnace is distinctive in design. A fire-brick arch is sprung over the 
grates immediately in front of the first bank of tubes. The large tri- 
angular space between boiler front, tubes, and mud drum forms the 
combustion chamber. 



138 



STEAM POWER PLANT ENGINEERING 




Fl«. 57. 2365-horsepower Stirling Boiler — Delray Station, Detroit Edison Company. 



BOILERS 139 

Fig. 57 gives a sectional view through the boiler and setting of a 
2365-horsepower Stirling boiler equipped with Taylor stokers as in- 
stalled at the Delray station of the Detroit Edison Company. Five 
boilers are now in operation and it is planned to eventually install ten. 
Though rated at 2365 boiler horsepower they are capable of carrying 
continuously a load equivalent to 6000 kilowatts with a maximum of 
8000 kilowatts. The overall dimensions of the boiler and setting are 
shown in the illustration. Each unit contains 23,654 square feet of 
effective heating surface and is provided with superheaters for supplying 
steam at 150 degrees superheat. Table 33 gives a resume of the prin- 
cipal results obtained from tests of these units with Roney and Taylor 
stokers. The grate surface per boiler for the Roney stoker is 446 
square feet and for the Taylor stoker 405 square feet, thus giving as 
ratios of grate surface to heating surface 1 : 53 and 1 : 58.5 respectively. 
For a complete description of these tests see Jour. A.S.M.E., Nov., 
1911, p. 1439. 

The largest boilers in this country (1917) are installed in the new 
Highland Park plant of the Ford Motor Company. Each unit contains 
25,000 sq. ft. of effective heating surface and furnishes 4000 boiler 
horsepower continuously. These boilers are of the Badenhausen type 
and are equipped with Taylor stokers (Power, Oct. 3, 1916, p. 474). 

71. Winslow High-pressure Boiler. — The standard types of boiler de- 
scribed in the preceding paragraph are seldom designed for pressure 
exceeding 250 lb. per sq. in. A few installations have been made for 
working pressures as high as 350 lb. per sq. in., but it is doubtful if this 
pressure will be exceeded without considerable modification in basic 
design. The weak element lies in the drum since excessive thickness 
of material is necessary for pressures above the limit mentioned. For 
example, the 60-in. drums of the Babcock & Wilcox boilers for the 
Joliet plant of the Public Service Company of Northern Ilhnois are If 
in. thick. With the prospect of pressures ranging as high as 1000 lb. 
per sq. in. (see paragraph 179) engineers are interested in types of boilers 
which can be built commercially to withstand these high pressures. 
Fig. 58 shows a section through the setting and one element of a ''Win- 
slow Safety High-pressure" boiler which may be designed to with- 
stand working pressures considerably in excess of 1000 lb. per sq. in. 
The assembled boiler consists of a number of sections, similar to the 
one illustrated in Fig. 58, forming a closely nested mass of tubes, each 
section being connected to a common steam header, feed pipe and mud 
drum. Referring to the illustration: each section is composed of a 
''front section header" A and "rear section header" B, connected by 
a number of approximately horizontal tubes, C, all made of seamless 



140 



STEAM POWER PLANT ENGINEERING 



steel tubing. The lower tubes are inclined, the front ends being higher 
than the rear. This degree of inclination gradually decreases in the 
upper tubes until the highest tube is practically horizontal. All tubes 
are slightly curved, the lower ones more than the upper. This preserves 




Fig. 58. Winslow "Safety High-pressure" Boiler and Hand-fired Setting. 



each tube in its original plane, even if it should expand considerably 
under heat. 

All joints between the tubes and section headers are welded. Extra 
material is added in the welding process and the joint is thus made 
stronger than the tube itself. 

Each section carries a baffle D, riveted in place, in contact with 
each side of the section, each baffle touching the one on the adjoining 
section, the outside ones being in contact with the wall of the enclosure. 
These several baffles form a complete baffle wall, which is in contact 
with the fire bridge E, confining the first and most intense action of 
the fire to the front part of the section. These baffles are either made 
of cast iron or of steel channels fifled with plastic refractory material. 



BOILERS 



141 



Connection to 
Steam Header 



Weld 



Water Level. 



Glass 
/Tube 



The baffles being in metallic contact with the tubes, their temperature 
can never greatly exceed that of the water or steam. 

All sections are supplied with water from a common feed pipe F 
closed at one end and carrying the check valve and pump connection 
at the other. A branch tube leads to each section entering the rear 
section header somewhat above its lower end. The joints at header 
and feed pipe are clearly shown. 

From the upper end of each rear section header a steel tube leads 
to the steam header G. This tube consists of two parts, one welded to 
the steam header and the other to the rear section header, connected 
by a special joint. To insure equal distribution of the flow of steam over 
the length of the steam header, the steam is taken through a large number 
of small holes, properly distributed, in a tube located inside the steam 
header and passing through one of its sealed ends. 
The lower end of the rear section header is 
formed into a special joint, which is connected to 
the mud drum H. The opening into the mud 
drum is as large as it is possible to make it and 
the passage is straight and without obstructions. 
Connected to the mud drum is the blow-off valve /, 
shown in dotted lines, Fig. 58. 

The three unions on each section at steam 
header, feed pipe and mud drum, are the only 
joints which are not welded, but these are all 
located in the last pass of the furnace gases and 
are not subjected to high temperatures. They 
are easily accessible and are made with metal to 
metal contact, without gaskets or packing. 

At high pressures and temperatures the ordi- 
nary gauge glasses are not desirable. One of the 
best indicators for severe conditions is shown in 
section in Fig. 59. The water column is a steel 
tube, surrounded by a steel jacket, the space be- 
tween being filled with mercury, visible in a ver- 
tical glass tube. That part of the mercury which 
surrounds steam in the column absorbs much 
more heat than the part which surrounds water, 
perature of the mercury and consequently its height in the glass tube 
is, therefore, a positive indication of the water level, being high for 
low water and low for high water. There is, of course, a certain 
lag in the indication on account of the time necessar>^ to transfer 
the heat through the metal wall of the column, but this is negligible 



^Mercury 



Connection to 
Mud Drum 



-Weld 



Fig. 59. Water Level 
Gauge — Winslow 
High-pressure Boiler. 

The average tem- 



142 STEAM POWER PLANT ENGINEERING 

on account of the wide range of water content which is permissible 
in the Winslow boiler without danger or improper influence on its 
operation. When superheated steam is produced a thermometer can 
be used as an additional indicator of the equivalent water level. 
This instrument is usually of the dial form, its bulb being inserted into 
the steam pipe at the flange connecting it to the steam collector. The 
dial, being connected to the bulb by a flexible tube, can be located at 
any point most convenient to the operator. A drop in temperature 
indicates a higher water level and vice versa. 

Circulation is as follows: When heat is applied to the boiler its 
first effect is to expand the water which is contained in that part of 
the section forward of the baffle, thereby reducing its specific gravity. 
Expansion causes the water column to rise in front and the heated water 
to flow toward the rear in the upper tubes. Reduced specific gravity 
in the front part causes a forward flow in the lower tubes. 

Each particle of water absorbs a certain amount of heat during cir- 
cuit through the set of tubes. When its temperature has reached 212 
deg. fahr. further absorption of heat causes the generation of steam, 
or, in other words, a sudden increase of volume and a consequent 
reduction of the specific gravity of the water. The immediate effect 
of this is to make the circulation more active. The rising column in 
the front section header then consists of a mixture of steam and water. 

The water drains back toward the rear section header through the 
upper circulation tubes, and the steam naturally tends to separate from 
the water at this point and to flow through the front header and the 
top tubes, toward the point of discharge. The returning water does not 
completely fill the upper tubes of the ''zone of circulation and evapora- 
tion, " but it exposes a certain amount of surface, from which further 
separation takes place of such steam as is carried along with the water 
or as is generated within the return fiow tubes. The foregoing will 
make it clear that the office of the upper part of the circulating tubes, 
which have been designated as ''return flow tubes," is to intercept 
the rising column of water and steam in the front section header, carry 
the water back by gravity, and prevent its entering the uppermost 
tubes. It should be noted that the inclination of these return flow 
tubes gradually decreases toward the top, as the amount of water they 
carry becomes less. 

The uppermost tubes practically contain steam only, and, being lo- 
cated in the flow of the hot gases, they effectively dry the steam. It is 
even possible, without any further provision, to superheat the steam 
somewhat in this "drying zone" before it is discharged from the section. 
If a higher degree of superheat is desired, the separate flue L, already 



BOILERS 143 

referred to and shown in Fig. 58 is provided. It carries the desired 
amount of hot gases from the furnace, in front of the nest of tubes, 
directly over the top of the boiler, through the ''drying zone," which 
thereby becomes a "superheating zone." 

72. Unit of Evaporation. — The performance of a boiler and furnace 
is commonly expressed in terms of the weight of water evaporated per 
pound of fuel or of the weight evaporated per hour per square foot of 
heating surface. To reduce all performances to an equal basis so as to 
facilitate comparison the evaporation under actual conditions is con- 
veniently referred to the equivalent evaporation from a feed water 
temperature of 212 deg. fahr. to steam at atmospheric pressure. The 
heat required to evaporate one pound of feed water at a temperature of 212 
deg. fahr. into saturated steam of the same temperature, or from and at 
212 deg. fahr. as it is conamonly called, is designated as one unit of evapo- 
ration (U.E.). The 1915 A.S.M.E. Boiler Code stipulates the use of 
Mark's and Davis' value for the latent heat of steam at 212 deg. fahr. 
and defines the standard unit of evaporation as 970.4 B.t.u. G. A. 
Goodenough (Properties of Steam and Ammonia, 1915, John Wiley & 
Sons, Publishers) assigns a value of 971.7 to this quantity and intimates 
that the correct value may be even slightly greater. The ratio of the 
heat necessary to evaporate one pound of water under actual con- 
ditions of feed temperature and steam pressure and quality to the heat 
required to evaporate one pound from and at 212 deg. fahr. is called 
the factor of evaporation. Thus for dry saturated steam, using Mark's 
and Davis' value for the latent heat, 

in which 

F = factor of evaporation, 

X = total heat of one pound of steam at observed pressure above 

32 deg. fahr., 
^2 = total heat of one pound of feed water above 32 deg. fahr. 

If the steam is wet, 

\ = xr-\-q, (31) 

in which 

X = the quality of the steam, 

r = latent heat of evaporation at observed pressure, 
q = heat in hquid at observed pressure. 
If the steam is superheated, 

X = r + g + Ct„ (32) 

* For most purposes qz may be taken at ^ — 32, in which ti = temperature of 
the feed water, deg. fahr. 



144 



STEAM POWER PLANT ENGINEERING 



in which 

C = the mean specific heat of the superheated steam, 
ts = the degree of superheat, deg. fahr. 

73. Heat Transmission. — Fig. 60 shows a section through a boiler- 
heating plate and serves to illustrate the accepted theory of heat trans- 
mission. The outer surface of the plate is covered with a thin layer 



Dry guriacev, 




wetsurface of soot and a film of gas, and the 

inner surface is similarly protected 
by a layer of scale and a film of 
steam and water. It is, therefore, 
reasonable to assume that the dry 
surface of the plate is located some- 
where within the film of gas, and 
the wet surface within the film of 
water and steam. 

The heat is imparted to the dry 
surface by: (1) radiation from the 
hot fuel bed and furnace walls, and 
by (2) convection from the moving 
furnace gases. The heat is trans- 
ferred through the boiler plate and 
its coatings purely by conduction. 
The final transfer from the wet sur- 
face to the water is mainly by 
convection. 

Radiation depends on the tem- 
perature, and according to the law 
of Stefan and Boltzmann is approximately proportional to the difference 
between the fourth power of the absolute temperature of the fuel bed 
and furnace walls and the temperature of the dry surface of the heating 
plate. According to this law the heat transmitted by radiation in- 
creases rapidly with the increase in furnace temperature. In the or- 
dinary boiler and setting the surface exposed to radiation is only a 
small portion of the total heating surface, and, since in well-operated 
furnaces the temperature of the furnace cannot be increased materially 
on account of practical considerations, there is little hope of increas- 
ing the capacity of a boiler by increasing the furnace temperature. 
The extent of heating surface exposed to radiation, however, may be 
greatly increased. Many authorities are of the opinion that the boiler 
of the future will depend largely upon radiation. That this predic- 
tion is being reahzed is evidenced by the high combustion efficiency 



A = Average Temperature of- Moving Gases. 
B= Average Temperature of Dry Surface. 
C = Average Temperature of Wet Surface. 
D =Temperature of Water in Boiler. 

Fig. 60. Heat Transmission through 
Boiler Plate. 



BOILERS 145 

and extremely high ratings effected by the modern duplex furnace, 
Figs. 57 and 103, in which a considerable portion of the boiler heating 
surface is exposed to direct radiation. 

The amount of heat imparted by convection from heated gases to 
cooler metal surfaces has been the subject of a great deal of investi- 
gation both from the experimental and theoretical side. Numerous 
attempts have been made to correlate the experimental data with the 
theoretical deductions but the results have been far from harmonious. 
This, however, has had little effect on the practical development of 
the boiler and it is quite probable that a more complete understanding 
of the phenomena will have no radical effect on the present design. 

The resistance of the metal itself is so small that it may be neglected 
in calculating the total heat transmission and it may be logically as- 
sumed that the plate will take care of all the heat that reaches its dry 
surface. 

The three distinct methods of heat transfer, radiation, convection 
and conduction, do not exist separately in the modern steam boiler 
but are operating at the same time. For this reason and in view of 
the number of arbitrary coefficients entering into the theoretical treat- 
ment of each method of heat transfer, engineers find it simpler to con- 
sider only the total heat transfer and to use empirical or semi-empirical 
equations. Thus, the total heat transfer, assuming no losses in the 
transmission, may be expressed 
in which SUd^WC^^, (33) 

S = square feet of heating surface, 

U = mean coefficient of heat transfer, B.t.u. per sq. ft. per degree 

difference in temperature per hour, 
d = mean temperature difference between the heated gases and the 
metal surface, deg. fahr., 
W = weight of gases flowing, lb. per hour, 
Cp = average mean specific heat of the gases, 
tm = mean temperature drop of the gases between furnace and 

breeching, deg. fahr. 
In practice U varies from 30 or more in the first row of tubes of a 
water tube boiler directly over the incandescent fuel bed to 5 or less 
in the last row immediately adjacent to the uptake. 

Experiments conducted by Jordan and the Babcock & Wilcox Com- 
pany indicate that the value of U varies approximately as follows: 

U = K + B ^* (34) 

* Trans. Int. Eng. Congress, "Mechanical Engineering," 1915, p. 366. 



146 STEAM POWER PLANT ENGINEERING 

in which 

K = coefficient determined experimentally, 

B = 2i function of the dimension of air passage and mean tempera- 
ture difference of the gas and metal, 
A = average cross sectional area of the gas passages through the 
boiler. 

Other notations as in equation (33) . 

For the standard type of Babcock & Wilcox water tube boiler, the 
Company's investigators found the following modification of equation 
(34) to give satisfactory results for 100 to 150 per cent ratings. 

^ = 2.0 + 0.0014^. (35) 

The curves in Fig. 61 may be used as a guide in approximating the 
heat transfer in fire tube boilers. 

An examination of equation (34) shows that for a given set of con- 
ditions and within certain limits the rate of heat transfer varies directly 
with the weight of gases flowing per unit area of gas passage. This is 
not strictly true since the rate of heat transfer varies as some power of 
the weight less than unity. But within narrow limits it is sufficiently 
accurate to consider the exponent as unity. 

Experiments by Professor Nicholson* and the U. S. Geological 
Survey t show that by establishing a powerful scrubbing action between 
the gases and the boiler plate the protecting film of gas is torn off as 
rapidly as it is formed and new portions of the hot gases are brought 
into contact with the plate, thereby greatly increasing the rate of heat 
transmission. Similarly, the faster the circulation of the water the 
greater will be the scrubbing action tending to remove the bubbles 
of steam from the wet surface and the more rapid will be the transfer 
from the plate to the boiler water. 

Professor Nicholson found that by filling up the flue of a Cornish 
boiler with an internal water vessel, leaving an annular space of only 
1 inch around the latter, an evaporation eight times the ordinary rate 
was effected at a flow of gases 330 feet per second (8 to 10 times the 
average flow). The fan for creating the draft consumed about 4 J per 
cent of the total power. 

The conclusion is that the heating surface for a given evaporation 
at the present rating may be reduced as much as 90 per cent for the same 
output, with a corresponding reduction in the size, cost, and space 
requirements, or with a given heating surface of standard rating the 

* Proc. Inst, of Engr. & Shipbuilders, 1910. 
t Bui. 18, U. S. Bureau of Mines, 1912. 



BOILERS 



147 



output may be enormously increased; also the increase in power 
necessary to create the draft is by no means comparable with the ad- 
vantages gained. 

The modern locomotive boiler is the nearest approach to these con- 
ditions in practice. Here a powerful draft forces the heated gases 



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OLb. 



2000 4000 6000 8000 10000 12000 

Weight of Gases per Sq. Ft. of Flue Area per Hour 



UOOOLb. 



Fig. 61. 



Heat Transfer in Boiler Flues. Results of Experiments by the 
Babcock & Wilcox Company. 



through small tubes at a very high velocity and an enormous evapo- 
ration is effected with a comparatively small heating surface. 

These principles have been applied to a limited extent to stationary 
boilers already installed by making the gas passages smaller as com- 



148 STEAM POWER PLANT ENGINEERING 

pared to the length by means of suitable baffles (Fig. 53) and by forc- 
ing larger weight of gas through the boiler, either by forced draft or 
by increasing the grate area (Fig. 103). 

In a general sense when the capacity of a boiler is doubled or tripled 
the over-all efficiency of the whole steam-generating apparatus drops, 
but the advantage gained usually offsets the loss in fuel economy. 
In the modern central station equipped with mechanical stokers of the 
forced draft type the boilers are operated normally on a basis of 5 to 
6 sq. ft. of heating surface per boiler horsepower, and at peak loads 
3 sq. ft. per horsepower is not unusual. 

On the Transmission of Heat in Boilers, Hedrick & Fessenden, Jour. A.S.M.E., 
Aug. 1916, p. 619. 

Heat Transmission in Boilers, Kreisinger and Ray: Tech. Paper 114, U. S. Bureau 
of Mines, 1915; Power and Engr., June 29, 1909, p. 1144; Bulletin No. 18, U. S. 
Bureau of Mines, 1912. 

Some Notes on Heat Transmission and Efficiencies of Boilers, R. Royds, Trans. 
Inst, of Engr. & Shipbuilders in Scotland, Vol. 58, 1915. _ 

The Heat of Fuels and Furnace Efficiency: W. D. Ennis, Power and Engr., July 
14, 1908, p. 50. 

A Study in Heat Transmission: (The transmission of Heat to Water in Tubes as 
Affected by the Velocity of the Water), J. K. Clement and C. M. Garland, Univ. 
of 111. Bulletin No. 40, Sept. 27, 1909; Power & Engr., Feb. 7, 1911, p. 222. 

Heat Transmission in Tubes, Dr. Wilhelm Nusselt: Zeit. d. ver. Deut. Ingr., 
1909, p. 750. 

On the Rate of Heat Transmission Between Fluids and Metal Surfaces, H. P. Jordan, 
Pro. Inst. Mec. Engrs., 1909. 

74. Heating Surface. — All parts of the boiler shell, flues, or tubes 
which are covered by water and exposed to hot gases constitute the 
heating surface. Any surface having steam on one side and exposed 
to hot gases on the other is superheating surface. According to the 
recommendations of the American Society of Mechanical Engineers, 
the side next to the gases is to be used in measuring the extent of the 
heating surface. Thus measurements are made of the inside area of 
fire tubes and the outside area of water tubes. The heating surface in 
a boiler under average conditions of good practice is most efficient when 
the heated gases leave the uptake at a temperature of 75 to 150 deg-. 
fahr. above that of the steam. Each square foot of heating surface is 
capable of transmitting a certain amount of heat, depending upon the 
conductivity of the material, the character of the surface, the temperature 
difference between the gases and the metal surface, the location and ar- 
rangement of the tubes and the density and the velocity of the gases. 

Thus one square foot of heating surface in the first pass of a water- 
tube boiler immediately over the incandescent mass of fuel may evapo- 
rate as high as 75 pounds of water per hour from and at 212 deg. fahr., 



I 



BOILERS 



149 



whereas the same extent of surface close to the breeching evaporates 
less than one pound per hour. Because of this extreme variation it is 
convenient to assume a uniform heat transmission for the entire surface 
which will give the same total evaporation as that actually obtained. 
For maximum economy under average conditions of hand fired operation 
this gives a meati evaporation of 3 to 3.5 pounds of water per square 
foot per hour from and at 212 deg. fahr., which is equivalent to allow- 
ing 10 to 12 square feet per boiler horsepower. By providing a large 
combustion chamber, increasing the extent of the first pass or the equiv- 
alent and by carrying a very thick bed of fuel a mean evaporation of 
10 pounds per square foot per hour has been maintained with high 
economy. This corresponds to 3.5 square feet of heating surface per 
boiler horsepower. 

The maximum evaporation is limited only hy the amount of coal which 
can he burned upon the grate. For example, a mean evaporation as high 
as 23.3 pounds * per square foot per hour has been effected in locomotive 
work, under intense forced draft, and 20 pounds per square foot per 
hour is not unusual in torpedo boat practice. Such extreme, high 
rates of evaporation, however, are invariably obtained at the expense 
of fuel economy. In the very latest central stations the boiler and 
settings are proportioned to operate at 100 per cent above standard 
rating with high over-all efficiency and at 200 per cent above rating 
with only a small drop in efficiency, but such results are not obtain- 
able in the ordinary hand fired boiler and setting. 

Builders of return tubular and vertical fire-tube boilers allow from 1 1 
to 12 square feet of heating surface per horsepower; water-tube boilers 
are rated at 10 square feet per horsepower, and Scotch-marine boilers 
at 8 square feet per horsepower. 

See, also, paragraph 79, Effect of Capacity on Efficiency. 

The following table shows approximately the relation between boiler 
horsepower and heating surface for different rates of evaporation : 

EVAPORATION FROM AND AT 212 DEG. FAHR. PER SQUARE FOOT PER HOUR. 



2 


2.5 


3.0 


3.5 


4 


5 


6 


7 


8 


9 


10 


SQUARE FEET HEATING SURFACE REQUIRED PER HORSEPOWER. 


17.3 


13.8 


11.5 


9.8 8.6 


6.8 


5.8 4.9 


4.3 


3.8 


3.5 



Efficiency of Boiler Heating Surface: Trans. A.S.M.E., 18-328, 19-571. Kent, 
"Steam Boiler Economy" (John Wiley & Sons), Chapter IX. 

The Nature of True Boiler Efficiency: Jour. West. Soc. Engrs., Sept. 18, 1907. 
Heat Tran^erence through Heating Surface: Engineering, 77-1. 

* Jour. A.S.M.E., Jan., 1915, p. 22. 



150 STEAM POWER PLANT ENGINEERING 

75. The Horsepower of a Boiler.* — A boiler horsepower is equivalent 
to the evaporation of 34.5 pounds of water per hour from a temper- 
ature of 212 deg. fahr, to steam at atmospheric pressure. This corre- 
sponds to 33,479 B.t.u. per hour. Since the power from steam is 
developed in the engine and the boiler itself does no work, the above 
measure of capacity is merely conventional but unfortunately leads to 
much confusion. Thus one boiler horsepower will furnish sufficient 
steam to develop about four actual horsepower in the best compound 
condensing engine, but only one-half horsepower in a small non-con- 
densing engine. Boilers should be purchased on the basis of heating 
surface and not on the horsepower rating, since one bidder may offer 
a boiler with, say, 5 square feet of heating surface per horsepower and 
another with 10 square feet, both being capable of the required evapo- 
ration, but the one with the small heating surface (which will, of course, 
be the cheaper boiler) will have considerably less reserve capacity. 
Manufacturers ordinarily rate their boilers on the basis of from 10 to 12 
square feet of heating surface per horsepower, and the power assigned 
is called the huilder^s rating. As this practice is not uniform, bids 
and contracts should always specify the amount of heating surface 
to be furnished. According to the recommendations of the American 
Society of Mechanical Engineers, ''A boiler rated at any stated capacity 
should develop that capacity when using the best coal ordinarily sold 
in the market where the boiler is located, when fired by an ordinary 
fireman, without forcing the fires, while exhibiting good economy. 
And, further, the boiler should develop at least one-third more than 
stated capacity when using the same fuel and operated by the same 
fireman, the full draft being employed and the fires being crowded; 
the available draft at the damper, unless otherwise understood, being 
not less than one-half -inch water column." 

In determining the boiler horsepower required for a given engine 
horsepower it is convenient to estimate the steam consumption of the 
engine under actual conditions and then ascertain the equivalent 
evaporation from and at 212 deg. fahr. For example, assume a simple 
non-condensing engine developing 20 horsepower to use 50 pounds of 
steam per horsepower hour, or 1000 pounds steam per hour; steam 
pressure, 80 pounds per square inch; feed-water temperature 120 deg. 
fahr. Required the boiler horsepower necessary to furnish this quantity 
of steam. 

* The unit "myriawatt" has been suggested by H. G. Stott as a unit of boiler 
capacity. For the conversion of myriawatts to other engineering units see Appen- 
dix F, 



BOILERS 151 

From equation (30), the factor of evaporation is 

^ X-^2 ^ 1185.3 - 87.91 ^ 

970.4 970.4 ^'^'^ ' 

One thousand pounds of steam under the given conditions are there- 
fore equivalent to 1000 X 1.131 = 1131 pounds from and at 212 deg. 
fahr. 

The boiler horsepower necessary to furnish steam for the 20-horse- 
power engine will be 

1131 
Boiler horsepower = ^^p^ = 32.8. 

Example 13. A 15,000-kilowatt steam turbine and auxiliaries require 
14.7 pounds of steam per kilowatt-hour at rated load; steam pressure, 
200 pounds per square-inch gauge; superheat, 150 deg. fahr.; feed- 
water temperature, 179 deg. fahr. 

Required the boiler horsepower necessary to furnish this quantity 
of steam. 

The heat furnished to the turbine and auxiliaries per kilowatt-hour is 

w{\ + Cpts - ^2) = 14.7 (1199.2 + 0.57 X 150 - 146.88) 
= 16,724 B.t.u., 

15 -1 u l^'OOO X 16,724 „__ , - 

Boiler horsepower = ^^^Tatg ^ '^^^ (approx.). 

If the boilers are to be operated at builder's rating (10 sq. ft. of heating 
surface per boiler horsepower) 75,000 sq. ft. of heating surface would 
be necessary. In plants of this size, however, the boilers would in all 
probability be operated at 200 per cent rating or more when furnishing 
steam at full load requirements, and 37,500 sq. ft. of heating surface 
would suffice. Assuming 200 per cent, the ratio of installed horsepower 
(builder's rating) to kilowatts of rated turbine capacity would be 1 to 4 
in this case. In large, modern central stations this ratio ranges between 
1 to 5 and 1 to 8 (reserve boilers not included). 

76. Grate Surface and Rate of Combustion. — The amount of fuel 
which can be burned per hour limits the amount of water evaporated 
per unit of time and depends upon the extent and nature of the grate 
surface, the character of the fuel and the draft. In locomotive and 
torpedo-boat practice space limitations necessitate the use of small 
grates and the rate of combustion is primarily a direct function of the 
draft. In stationary practice there is a wide permissible range in pro- 
portioning the grate surface, since a given rate of combustion may be 
effected with large grate surface and light draft or with small grate 
surface and strong draft. In a general sense the best results are obtained 
with a small grate and a high rate of combustion, but in the majority of 
installations the draft is comparatively feeble and a liberal grate area 
is necessary. So much depends upon the grade and size of the fuel 



152 



STEAM POWER PLANT ENGINEERING 



that general rules for proportioning the grate surface are apt to lead to 
serious error. A liberal allowance of grate surface is desirable for hand- 
fired furnaces with natural draft, particularly if the ash is easily fusible, 
tending to choke the grate, but with forced draft and automatic stokers 
the best results are obtained with a thick fire and small grate surface. 
The maximum grate area is limited by the dimensions of the furnace. 
In hand-fired furnaces with stationary grates the width of the furnace is 
limited by that of the boiler and the length by the distance which the 
fireman can control the fuel bed. Anthracite requires no slicing and a 
much greater length of grate can be manipulated than with the caking 
variety of coals. Shaking and self-dumping grates may be of greater 
length than stationary grates since hand manipulation is largely dis- 
pensed with. The dimensions of mechanical stokers depend largely 
upon the type of stoking device. In practice the maximum rate of 
combustion, pounds per square foot of grate surface per hour, is usually 
assumed and the grate area and chimney height or equivalent propor- 
tioned to effect the desired rate of combustion. See Table 29. 

The ratio of grate area to heating surface is sometimes used as a 
guide in proportioning the grate but the extent of grate surface depends 
upon so many other factors that this method of procedure is of little 
value and apt to lead to serious error. Thus, a study of several hundred 
boiler installations gave results as follows: 



Type of Grate or Stoker. 



Hand-fired . 
Hand-fired . 
Chain grate 

Roney 

Taylor 

Jones 



Kind of Coal. 



Anthracite 

Bituminous 

Bituminous 

Bituminous 

Bituminous 

Bituminous 



Ratio Grate Surface to Heating Surface. 



Minimum. 



1 to 30 
1 to 40 
1 to 36 
1 to 30=* 
1 to 50 
1 to 55 



Maximum. 



1 to 65 
1 to78 
1 to 72 
1 to 55 
1 to 82 
1 to68 



Double Stoker. 



A number of boiler tests made by Barrus (''Boiler Tests") showed 
that the best economy with anthracite coal, hand-fired, was obtained 
with an average ratio of grate surface to heating surface of 1 to 36, and 
at a rate of combustion of approximately 12 pounds of coal per square 
foot of grate surface per hour. In these tests a variation in grate and 
heating-surface ratio of 1 to 36 up to 1 to 46 gave practically no differ- 
ence in economy. With bituminous coal the tests showed that an 
average ratio of 1 to 45 gave the best results and at a rate of combustion 
of 24 pounds of coal per square foot of grate surface per hour. 



BOILERS 



153 



Tests made by Christie (Trans. A.S.M.E., 19-330) gave an average 
combustion of 13 pounds of anthracite per square foot of grate per hour 
for maximum efficiency and 24 pounds of bituminous. 

Table 32 gives the relation between heating and grate surface in a 
number of recent boiler installations using different kinds of coal, and 
is illustrative of current practice. 

The rate of combustion depends upon the grade and size of coal, 
thickness of fire, percentage of air spaces in the grate, available draft 
through the fire and the efficiency of combustion, and can only be found 
accurately by experiment. For a general set of conditions the rate of 
combustion is primarily a function of the pressure difference between 
the ash pit and furnace, and is approximately as shown in Table 29. 



TABLE 29. 

MAXIMUM ECONOMICAL RATE OF COMBUSTION. 

Pounds of Coal per Sq. Ft. of Grate Surface per Hour. 



Kind of Coal. 



Anthracite, No. 3 

Anthracite, No. 2 

Anthracite, Pea 

Semi-bituminous 

Ky., Pa., and Tenn. bitum 
111., Ind., and Kan. bitum. 





Force of Draft between Furnace and Ash Pit, 




Inches of Water. 


0.1 


0.15 


0.20 


0.25 


0.30 


0.35 


0.40 


0.45 


0.50 


3.5 


5.0 


6.2 


7.2 


8.2 


9.0 


10.0 


11.0 


12.0 


5.5 


7.2 


9.0 


10.5 


11.8 


13.0 


14.2 


15.5 


16.5 


8.0 


9.2 


11.5 


13.5 


15.0 


17.0 


18.5 


20.0 


21.5 


9.5 


13.0 


16.0 


19.0 


23.0 


25.0 


27.0 


30.0 


32.0 


10.0 


14.0 


18.0 


21.0 


24.0 


27.0 


30.0 


33.0 


36.0 


11.0 


15.0 


20.0 


24.0 


28.0 


32.0 


35.0 


38.0 


41.0 



With forced draft these rates of combustion may be greatly increased. 
Some idea of the extreme rate of combustion in modern locomotive 
practice may be obtained from the following figures which give the 
pounds of coal burned per hour per square foot of grate surface for 
various conditions of operation: 



Maximum rate 225 

Very high rate 150 

Average high rate 100 



Average rate 80 

Economical rate 60 

Low rate 50 



In proportioning the grate surface for a proposed installation the 
principal factor considered is the character of the fuel, a study being 
made of the various fuels available, and the one selected which gives 
the highest evaporation per dollar (all items entering into the handling 
and combustion of the fuel being considered). This information may 
usually be obtained from records of plants using the same grade of fuel 
and grates similar to those intended for the proposed plant. 



154 STEAM POWER PLANT ENGINEERING 

77. Boiler, Furnace and Grate Efficiency. — A perfect boiler and furnace 
is one which transmits to the water in the boiler the total heat of 
the fuel. In order to effect this result combustion must be complete, 
there must be no radiation or leakage losses and the products of com- 
bustion must be discharged at the initial temperature of the fuel. No 
commercial form of steam boiler can fulfill these conditions, hence the 
amount of heat absorbed by the boiler will always be less than the calo- 
rific value of the fuel. 

The efficiencies recommended by the A.S.M.E., Rules for Conducting 

Boiler Tests, 1915, may be expressed as 

Heat absorbed by the boiler 

^rn . fi M r IX per pound of coa/ as /ired ,^„. 

hithciencv oi boiler, lurnace and grate = ttH — ^ i ? ' , , (ob) 

Calormc value 01 one pound 

of coal as fired 
Heat absorbed by the boiler per pound 

^rr, • 11 1 x-i 1 oi combustibleburned on the grsite* .„_. 

Efficiency based on combustible = -p^r-. — r^ -. j ^-j — .• (37) 

Calorific value of one pound ot 

combustible as fired 

Example 14. Calculate the various boiler efficiencies from the fol- 
lowing data : 

DATA AS OBSERVED. 

Steam pressure, pounds per square inch (gauge) 151 .0 

Barometer, inches of mercury 28 . 5 

Temperature of feed water, deg. fahr 161 . 8 

Temperature of the furnace, deg. fahr 2100.0 

Temperature of flue gases, deg. fahr 480.0 

Temperature of boiler room, deg. fahr 60 . 

Quality of steam, per cent 98 . 

Water apparently evaporated, pounds per hour 86,000 

Coal as fired, pounds per hour 10,000 

Refuse removed from ash pit, pounds per hour 1600 

COAL ANALYSIS, PER CENT OF COAL AS FIRED. 

Moisture 8 

Ash 12 

B.t.u. per pound, 11,250. 

CALCULATED DATA. 
Water apparently evaporated per pound of coal as fired, pounds = 86,000 -^ 10,000 

= 8.60. 
Factor of evaporation f = [0.98 X 856.8 + 338.2 - (161.8 - 32)] -h 970.4 = 1.08. 

* The combustible burned on the grate is determined by subtracting from the 
weight of coal supplied to the boilers, the moisture in the coal, the weight of ash 
and unburned coal withdrawn from the furnace and ash pit and the weight of dust, 
soot, and refuse, if any, withdrawn from the tubes, flues and combustion chambers, 
including soot and ash carried away in the gases. 

t See footnote, par. 72. 



BOILERS 155 

Equivalent evaporation per pound of coal as fired, pounds = 8.6 X 1.08 = 9.288. 
Heat absorbed by the boiler per pound of coal as fired, B.t.u. = 9.288 X 970.4 = 

9,013.0. 
Efficiency of boiler furnace and grate, per cent = (9.013 -^ 11,250) 100 = 80.11. 
Refuse in ash referred to coal as fired, per cent = (1600 -^ 10,000) 100 = 16.0. 
Combustible burned on the grate, per cent of coal as fired = 100 — (8 + 16) = 76.0. 
Equivalent evaporation per pound of combustible burned, pounds = 9.288 -r 0.76 

= 12.221. 
Heat absorbed per pound of combustible burned, B.t.u. = 12.221 X 970.4 = 11,860. 

Combustible as fired, per cent = 100 - (8 + 12) = 80.00. 
Calorific value of the combustible as fired, B.t.u. = 11,250 -h 0.80 = 14,062. 
Efficiency based on combustible, per cent = (11,860 -^ 14,062) 100 = 84.34. 

For oil fuel furnaces and coal furnaces equipped with stokers and 
forced draft appliances the net efficiency of the boiler and furnace may 
be taken as the boiler and furnace efficiency minus the equivalent heat 
required to feed the fuel and to create the draft. 

Attempts have been made to separate the combined efficiency of 
boiler, furnace, and grate into two parts, viz., efficiency of the boiler 
alone and efficiency of the furnace and grate, but the results have been 
discordant and involve the use of factors which cannot be obtained with 
any degree of accuracy. Thus ''true" boiler efficiency has been defined 
as the ratio of the heat absorbed to that available. The ''heat ab- 
sorbed" is taken as the difference between the heat generated in the 
furnace and that discharged into the flue, and the "available" heat 
is defined as the difference between the heat generated in the furnace 
and that discharged by the products of combustion at the temper- 
ature of the saturated steam. 

If Wf, Wc = weight of the products of combustion in the furnace 

and passing through the uptake, respectively, lb, per 

hour 
Tf, Tc, Ts, T = absolute temperature of the furnace gases, flue gases, 

saturated steam and boiler room, respectively, deg. fahr. 
Cf, Cc, Cs = mean specific heat of the products of combustion for 

temperature ranges t to t/, tc, ts, respectively. 

Then, neglecting radiation and minor losses, the "true" boiler effi- 
ciency equals 

WfCfTf - WcCsTs 

Assuming no leakage, w/ = Wc] and neglecting the difference in the 
mean specific heats, Cf = Cc = c,. With these assumptions, equation 
(38) reduces to 



156 



STEAM POWER PLANT ENGINEERING 



TABLE 30. 

RELATION BETWEEN FUEL CONSUMPTION AND BOILER, FURNACE AND GRATE 

EFFICIENCY. 
(Pounds of Fuel Burned per Boiler Horsepower- hour.) 



Calorific Value 






Boiler, Furnace and Grate Efficiency 


. 




of Fuel, B.t.u. 
per Pound. 




















40 


45 


50 


55 


60 


65 


70 


75 


80 


85 


7,500 


11.17 


9.91 


8.94 


8.12 


7.45 


6.87 


6.37 


5.95 


5.58 


5.25 


8,000 


10.45 


9.30 


8.37 


7.60 


6.97 


6.43 


5.98 


5.58 


5.22 


4.92 


8,500 


9.84 


8.75 


7.87 


7.12 


6.56 


6.05 


5.62 


5.25 


4.97 


4.63 


9,000 


9.30 


8.25 


7.45 


6.76 


6.20 


5.72 


5.31 


4.96 


4.65 


4.36 


9,500 


8.80 


7.83 


7.05 


6.40 


5.87 


5.41 


5.02 


4.69 


4.40 


4.14 


10,000 


8.37 


7.44 


6.70 


6.09 


5.58 


5.15 


4.79 


4.46 


4.18 


3.94 


10,500 


7.98 


7.09 


6.39 


5.80 


5.86 


4.90 


4.56 


4.26 


3.99 


3.76 


11,000 


7.60 


6.79 


6.09 


5.52 


5.06 


4.67 


4.34 


4.05 


3.80 


3.59 


11,500 


7.28 


6.49 


5.83 


5.29 


4.85 


4.47 


4.16 


3.88 


3.64 


3.45 


12,000 


6.97 


6.22 


5.58 


5.06 


4.65 


4.28 


3.99 


3.72 


3.48 


3.28 


12,500 


6.69 


5.97 


5.35 


4.86 


4.46 


4.11 


3.82 


3.57 


3.34 


3.14 


13,000 


6.44 


5.74 


5.15 


4.68 


4.29 


3.96 


3.68 


3.43 


3.22 


3.02 


13,500 


6.20 


5.52 


4.96 


4.51 


4.18 


3.81 


3.54 


3.31 


3.10 


2.91 


14,000 


5.98 


5.33 


4.79 


4.35 


3.99 


3.68 


3.42 


3.19 


2.99 


2.81 


14,500 


5.77 


5.15 


4.62 


4.20 


3.84 


3.54 


3.30 


3.08 


2.88 


2.72 


15,000 


5.58 


4.96 


4.47 


4.06 


3.72 


3.43 


3.19 


2.98 


2.79 


2.64 



TABLE 31. 

RELATION BETWEEN RATE OF EVAPORATION PER POUND OF FUEL AND 

BOILER, FURNACE AND GRATE EFFICIENCY. 

(Pounds of Water Evaporated per Hour from and at 212 deg. fahr. per Pound of Fuel.) 



Calorific Value 






Boiler, Furnace and Grate Efficiencj 








of Fuel, B.t.u. 
per Pound. 






















40 


45 


50 


55 


60 


65 


70 


75 


80 


85 


7,500 


3.09 


3.48 


3.86 


4.25 


4.64 


5.02 


5.41 


5.80 


6.18 


6.57 


8,000 


3.30 


3.71 


4.12 


4.55 


4.95 


5.36 


5.77 


6.18 


6.60 


7.01 


8,500 


3.51 


3.94 


4.38 


4.81 


5.26 


5.70 


6.14 


6.57 


7.01 


7.45 


9.000 


3.71 


4.18 


4.64 


5.10 


5.56 


6.04 


6.50 


6.96 


7.42 


7.90 


9,500 


3.92 


4.41 


4.90 


5.39 


5.88 


6.47 


6.86 


7.35 


7.85 


8.33 


10,000 


4.12 


4.64 


5.16 


5.66 


6.19 


6.70 


7.21 


7.74 


8.25 


8.76 


10,500 


4.31 


4.86 


5.40 


5.94 


6.48 


7.01 


7.55 


8.10 


8.64 


9.17 


11,000 


4.52 


5.09 


5.65 


6.22 


6.79 


7.35 


7.91 


8.48 


9.05 


9.61 


11,500 


4.74 


5.31 


5.91 


6.50 


7.10 


7.69 


8.28 


8.86 


9.45 


10.0 


12,000 


4.94 


5.55 


6.16 


6.78 


7.40 


8.01 


8.64 


9.25 


9.86 


10.5 


12,500 


5.14 


5.78 


6.42 


7.06 


7.70 


8.35 


9.00 


9.64 


10.3 


11.0 


13,000 


5.35 


6.01 


6.69 


7.35 


8.01 


8.69 


9.35 


10.0 


10.7 


11.4 


13,500 


5.56 


6.25 


6.95 


7.65 


8.34 


9.03 


9.72 


10.4 


11.1 


11.8 


14,000 


5.75 


6.48 


7.20 


7.91 


8.64 


9.35 


10.1 


10.8 


11.6 


12.2 


14,500 


5.96 


6.70 


7.45 


8.20 


8.95 


9.70 


10.5 


11.2 


12.0 


12.7 


15,000 


6.18 


6.95 


7.72 


8.50 


9.26 


10.1 


11.8 


11.6 


12.4 


13.1 



BOILERS 157 

The maximum theoretical efficiency of the boiler or the efficiency 
of the ideal or perfect boiler, based on utilizing all the heat except the 
inherent losses, may be expressed as 

^2 = -^^—, (40) 

in which 
H = calorific value of the coal as ffi'ed, 
/ = inherent losses as analyzed in paragraph 36. 
The efficiency ratio or the extent to which the theoretical possibilities 
are realized may be taken as 

£3=J> (41) 

in which 

E = efficiency of the boiler, furnace and grate (A.S.M.E. code), 

E2 = as in equation (40). 

The furnace and grate efficiency based on heat available may be ex- 
pressed 

E^"^J^, (42) 

in which F = furnace losses consisting of the (a) loss due to unburned 
fuel dropping through the grate or withdrawn from the furnace, (6) loss 
due to the production of CO, (c) loss due to escape of unburned hydro- 
carbons, (d) loss due to the combustion of carbon and moisture and 
production of hydrogen when fresh moist coal is thrown on a bed of 
white hot coke, (e) radiation due to the furnace and (/) unaccounted for 
losses due to the furnace. (For an analysis of these losses see para- 
graphs 25 to 36.) 

Equation (42) does not furnish a method of finding the true effi- 
ciency because it is impossible to determine loss (d) and impracticable 
to obtain loss (c) with the gas testing appliances ordinarily available. 
It is also impossible to separate losses (e) and (/) attributed to the fur- 
nace from the boiler losses alone due to radiation and unaccounted for. 

In practice the operating engineer is chiefly concerned with the com- 
bined efficiency of the boiler, furnace and grate, as defined by the 
A.S.M.E. Boiler Code. This factor is readily determined with the 
ordinary instruments found in the average modern plant. Table 32, 
compiled from a number of tests of difTerent types of boilers with differ- 
ent kinds of stokers and grades of fuel, gives some idea of the range of 
efficiencies incident to general practice. In attempting to better the 
efficiency it is necessary to separate the various losses as described in 
paragraphs 25 to 35, since this procedure enables the engineer to lo- 
cate the source of loss, and by comparing the actual and inherent losses 



158 



STEAM POWER PLANT ENGINEERING 



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BOILERS 



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160 



STEAM POWER PLANT ENGINEERING 



show where improvement may be effected. Although efficiencies of 
80 per cent or more have been reaUzed in several instances without the 
use of economizers, such performances cannot be expected for con- 
tinuous operation. In pumping stations or in plants where there are 
no peak loads and the boiler may be operated under a constant set of 
conditions a continuous efficiency of 75 per cent has been realized 
with coal as fuel and 80 per cent with fuel oil, but these figures are ex- 
ceptional. In large central stations with the usual peak loads in the 
morning and evening and long banking periods, over-all yearly effi- 
ciency is seldom greater than 70 per cent, though the boilers may be 
giving 77 to 81 per cent efficiency when operating at the most economical 
load. In large isolated stations with variable loads an over-all boiler 
and furnace efficiency on the yearly basis of 65 per cent is exceptional 
and a fair average is not far from 60 per cent. Small stations that 
show at times an efficiency as high as 75 per cent seldom average 50 



TABLE 33. 

PRINCIPAL DATA AND RESULTS OF TESTS ON 2365-RATED-HORSEPOWER STIRLING 

BOILERS AT THE DELRAY STATION OF THE DETROIT EDISON COMPANY. 

Tests with Roney Stoker. R6suin6 of Principal Results. 



No. of 
Test. 



1 

2 

3 

4 

5 

6 

16 

17 

18 

2^t 

5-6 1 



Length, 
Hr. 



Per Cent 
Rating. 



105.0 

80.0 

113.8 

152.4 

94.0 

150.7 

98.6 

193.3 

195.7 

119.8 

127.3 



B.t.u. in 
Coal. 



14,362 
14,225 
14,308 
13,756 
13,896 
14,037 
14,476 
14,493 
13,689 
14,098 
13.977 



Per Cent 

Ash in 
Dry Coal. 



5.98 
6.52 
7.40 
6.54 
6.89 
6.13 
9.68 
8.24 
9.81 
6.81 
6.84 



Efficiency. 



77.84 
79.88 
77.45 
75.78 
81.15 
75.28 
80.98 
76.73 
75.57 
76.13 
76.23 



Per Cent 
Steam used 
by Stoker 

Engines 

and Steam 

Jets. 



0.63 
1.58 
1.75 
1.45 
1.34 
1.39 
1.32 



Per Cent 
Combusti- 
ble in Ash. 



17.9 
24.4 
30.8 
31.6 
26.7 
34.1 
24.6 
23.2 
25.8 
29.4 



Temp, of 

Flue Gases 

Leaving 

Boiler, 

Deg. Fahr. 



576 
480 
542 
670 
483 
662 
460 
636 
694 
572 
575 



t Including periods between tests. 

Tests with Taylor Stoker. R6sum6 of Principal Results. 



No. of 
Test. 


Length, 
Hr. 


Per Cent 
Ratmg. 


B.t.u. in 
Coal. 


Per Cent 

Ash in 
Dry Coal. 


Efficiency. 


Per Cent 
Steam used 

by Stoker 
Auxiliaries.* 


Per Cent 
Combusti- 
ble in Ash. 


Temp, of 
Flue Gases 

Leaving 

Boiler, 
Deg. Fahr. 


7 


24 


151.2 


14,000 


7.03 


77.07 


2.61 


31.5 


575 


8 


24 


107.9 


13,965 


6.34 


80.28 


2.44 


27.1 


493 


9 


50 


162.8 


13,998 


6.75 


77.85 


2.87 


31.3 


574 


10 


48 


92.9 


14,188 


9.90 


77.90 


2.63 


27.2 


487 


11 


26.5 


211.3 


14,061 


9.55 


75.84 


3 41 


36.1 


651 


12 


48 


121.3 


14,010 


8.09 


79.24 


2.57 


27.6 


535 


14 


24 


185.2 


14,272 


8.71 


76.42 


2.95 


28.8 


647 


15t 


24 


123.1 


14,213 


8.34 


74.90 


2.77 


30.1 


561 


7-9 1 


109 


140.0 


13,983 


7.22 


77.66 


2.68 


29.9 


545 


lo-m 


80.5 


132.8 


14,095 


9.58 


75.66 


3.04 


31.1 


542 



* Engines driving stokers and steam-turbine driving fan. 

t In test No. 15 the fires were banked for 7^ hours and the averages include this period. 

t Including periods between testa. 



BOILERS 



161 



per cent to the year. These figures refer to boiler installations without 
economizers. For influence of the latter on boiler, furnace and grate 
efficiency see Paragraph 285. In general the over-all efficiency is 
dependent primarily on the load factor. The greater the load factor 
the smaller will be the standby losses (see Paragraph 35) and the nearer 
will the over-all efficiency approach test results. The usual discrepancy 
between efficiency as determined by special tests and average operation 
is due to the fact that the efficiency test is usually conducted under 
ideal conditions. The boiler surfaces are cleaned, the rate of combustion 
carefully adjusted to maximum economy and special attention given the 



TABLE 34.* 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, " Standard " Setting.) 

Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 











Horse- 


Heat 


Total 
Heating 
Surface 
per Horse- 
power. 


Super- 


Dry Coal 


Test 


Date, 


Horse- 


Efficiency, 


power per 


Lost in 


heat of 


per Sq. Ft. 


No. 


1908. 


power. 


Per Cent. 


Sq. Ft. 
Grate. 


Refuse, 
Per Cent. 


Steam, 
Deg. Fahr. 


G. S. 
per Hour. 


2 


Mar. 9 


873 


67.4 


9.70 


2.8 


6.76 


197 


41.2 


4 


" 10 


873 


69.0 


9.52 


2.8 


6.89 


195 


39.1 


6 


" 11 


852 


67.3 


9.47 


2.8 


6.93 


189 


38.9 


8 


" 16 


836 


65.3 


9.29 


6.4 


7.06 


174 


39.5 


10 


" 17 


870 


68.8 


9.67 


5.0 


6.78 


180 


39.3 


14 


" 19 


920 


66.2 


10.22 


9.2 


6.42 


187 


43.7 


16 


" 23 


900 


69.5 


10.00 


4.0 


6.56 


181 


40.5 


18 


" 24 


916 


69.1 


10.18 


5.5 


6.44 


190 


41.6 


20 


" 26 


912 


69.2 


10.13 


4.4 


6.48 


179 


41.2 


22 


" 27 


906 


67.7 


10.07 


4.1 


6.52 


194 


42.5 


24 


" 30 


925 


69.8 


10.28 


2.8 


6.38 


179 


41.6 


26 


" 31 


894 


69.4 


9.93 


5.2 


6.60 


170 


40.6 


28 


Apr. 1 


922 


71.2 


10.24 


3.6 


6.40 


169 


40.4 


30 


" 2 


923 


71.5 


10.26 


4.6 


6.40 


173 


40.5 


32 


" 7 


914 


70.0 


10.20 


4.5 


6.46 


175 


40.9 


34 


" 8 


939 


73.8 


10.4 


3.8 


6.28 


181 


40.4 


36 


" 10 


911 


70.9 


10.1 


3.0 


6.48 


185 


40.2 


38 


" 11 


967 


70.1 


10.7 


3.0 


6.11 


192 


42.6 


40 


" 13 


995 


67.8 


11.1 


3.4 


5.93 


211 


43.6 


42 


" 14 


887 


66.8 


9.9 


4.5 


6.65 


202 


40.8 


44 


" 27 


880 


69.5 


9.8 


5.5 


6.72 


169 


39.7 


48 


" 29 


927 


71.5 


10.3 


3.3 


6.37 


171 


40.8 


50 


" 30 


899 


70.3 


10.0 


4.2 


6.57 


171 


39.6 


52 


May 6 


886 


69.4 


9.8 


5.3 


6.67 


171 


38.2 


54 


" 7 


900 


69.1 


10.0 


4.8 


6.56 


171 


39.2 


56 


" 8 


967 


71.9 


10.7 


4.8 


6.10 


164 


40.1 


58 


" 11 


902 


70.5 


10.0 


3.3 


6.55 


163 


39.6 


60 


" 13 


875 


70.7 


9.7 


3.8 


6.74 


147 


38.3 


64 


" 14 


1102 


72.0 


12.2 


4.8 


5.35 


180 


43.2 



* This unit is still (1916) in operation and while the baffling has been changed the results are approxi- 
mately the same as given in the table. 



162 



STEAM POWER PLANT ENGINEERING 



TABLE 34. (Continued.) 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, " Standard " Setting.) 

Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 



Draft 


B.t.u. per 


Ash in 


Ash in 


Uptake 


CO2, 


Heat Lost 
up Stack 


Over Fire. 


In Uptake. 


Pound 
Dry Coal. 


Dry Coal, 
Per Cent. 


Refuse, 
Per Cent. 


Temp. 
Deg Fahr. 


Per Cent. 


(Dry Gas), 
Per Cent. 


0.87 


1.34 
1.25 
1.25 


11,634 
11,759 
12,039 


18.46 

16.81 
16.08 


82.33 

81.36 
80.03 


466 
461 
463 


6.9 

6.7 

7.7 




0.78 




0.83 


""'l5;6'"' 


0.94 


1.34 


11,993 


15.91 


67.42 


477 


7.6 


16.8 


0.84 


1.24 


11,909 


15.71 


71.32 


475 


7.9 


16 2 


0.99 


1.41 


11,768 


16.04 


63.78 


479 


8.5 


15.4 


0.77 


1.17 


11,846 


16.68 


79.04 


483 


9.1 


14.0 


0.81 


1.25 


11,800 


16.39 


71.98 


484 


8.3 


15.8 


0.77 


1.21 


11,846 


15.51 


78.53 


486 


9.0 


14.5 


0.78 


1.22 


11,659 


17.59 


80.58 


494 


9.2 


14.6 


0.68 


1.28 


11,800 


16.22 


82.97 


487 


8.8 


15.1 


0.70 


1.24 


11,752 


16.18 


76.84 


484 


8.8 


15.1 


0.62 


1.21 


11,862 


15.38 


82.99 


480 


9.2 


14.1 


0.58 


1.40 


11,800 


16.02 


78.37 


480 


9.1 


14.4 


0.73 


1.24 


11,815 


16.84 


77.84 


494 


9.0 


14.7 


0.72 


1.25 


11,659 


18.06 


82.27 


504 


8.9 


15.3 


0.65 


1.13 


11,831 


17.15 


86.92 


493 


9.7 


13.4 


0.70 


1.24 


12,002 


16.05 


84.39 


502 


9.0 


15.1 


0.71 


1.23 


12,469 


14.87 


82.14 


522 


9.7 


13.3 


0.63 


1.09 


12,049 


15.17 


78.12 


500 


9.5 


13.3 


0.71 


1.26 


11,801 


15.75 


77.21 


470 


8.3 


15.7 


0.68 


1.23 


11,769 


18.59 


84.04 


472 


8.7 


14.2 


0.66 


1.27 


11,955 


16.11 


79.30 


473 


7.9 


16.1 


0.62 


1.20 


12,360 


13.63 


74.59 


476 


8.8 


14.5 


0.66 


1.31 


12,298 


13.62 


75.19 


480 


9.0 


14.4 


0.66 


1.29 


12,423 


13.37 


75.61 


474 


9.4 


13.3 


0.92 


1.18 


11,956 


17.45 


83.24 


451 


9.2 


12.5 


0.76 


0.98 


11,971 


17.45 


80.99 


443 


10.0 


11.2 


0.68 


1.15 


13,126 


10.24 


70.90 


487 


10.4 


12.1 



firing, whereas, in most plants these refinements are seldom attempted. 
In our strictly modern boiler plants, refinement of design and a syste- 
matic supervision of operation have resulted in over-all efficiencies far 
above anything hitherto thought possible. 

The boiler, furnace and grate efficiency is only one of the many 
factors entering into the economical operation of the boiler plant. 
Different fuels may give the same efficiency under actual operating 
conditions, but the ultimate economy in dollars and cents may vary 
considerably. The real criterion is the net cost of evaporation, taking 
into consideration the cost of handling the fuel, disposition of refuse, 
ability to handle peak loads and depreciation of grate and setting. 



BOILERS 163 

The common practice of comparing the performance of boilers on the 
"fuel cost to evaporate 1000 pounds of water," is apt to lead to erro- 
neous conclusions; thus, the cost of evaporating 1000 pounds of water 
from and at 212 deg. fahr., per pound of cheap bituminous screenings, 
miay be 12 cents as against 18 cents per pound of high-grade and more 
costly washed coal, but the freight charges, cost of handhng the fuel 
and disposition of the ash may more than offset the gain in evaporation 
and the cheaper fuel may prove to be the more expensive in the end. 
Each installation is a problem in itself and all local influencing con- 
ditions must be considered before maximum economy can be effected. 
In general, for plants equipped with coal and ash handling machinery 
and adjacent to a railroad or to water transportation, the cheaper the 
fuel per pound of combustible the lower will be the ultimate cost of 
evaporation. 

Report of the Power Test Committee, A.S.M.E. Boiler Code, 1915, Jour. A.S.M.E., 
Vol. 37, 1915, p. 1273. This report may be had in pamphlet form. See also Ap- 
pendix A. 

78. Boiler Capacity. — Boilers are ordinarily rated on a commercial 
basis of 10 square feet of heating surface per horsepower. This rating 
is absolutely arbitrary and implies nothing as to the limiting amount of 
water that this amount of heating surface will evaporate. It has long 
been known that the evaporative capacity of a well-designed boiler is 
limited only by the amount of fuel that can be burned on the grate. 
Thus in locomotive practice a boiler horsepower has been developed with 
two square feet of heating surface and in torpedo boat practice this 
figure has been reduced to 1.8 square feet. If there were no practical 
limitations to capacity few, if any, boilers would be operated at the 
rated load and the amount of heating surface for a given evaporation 
would be only a fraction of the present requirements. Briefly stated 
the limitations are: 

1. Efficiency. — As the capacity increases beyond a certain limit 
the over-all efficiency drops off and a point is reached where further 
increase in capacity is obtained at a cost greater than that of additional 
heating surface. 

2. Grate Surface. — All fuels have a maximum rate of combustion 
beyond which satisfactory results cannot be obtained. With this 
limit established the only method of obtaining added capacity is through 
the addition of grate surface. Since the grate surface for a given boiler 
is limited by the impracticability of operating economically above a 
certain size there is obviously a commercial limit to the maximum 
weight of fuel burned per unit of time. 



164 



STEAM POWER PLANT ENGINEERING 



3. Draft. — In order to effect a heavy rate of combustion a great 
increase in draft is necessary. Apart from the power required to pro- 
duce the draft there is the loss of fuel carried away in the '' cinders. '^ 

4. At heavy rates of driving the furnace and stoker upkeep may 
become excessive. 

5. Feed Water. — For continuous high boiler overloads the feed 
water must be practically free from scale-forming elements and matter 
which tends to cause foaming and priming. 

TABLE 35. 

RELATION BETWEEN CAPACITY AND EFFICIENCY. 
(Evaporation from and at 212 Deg. Fahr. per Square Foot of Heating Surface per Hour.) 



2 


2.5 


3 


3.5 


4 


5 


6 


8 


10 


12 


Probable Relative Economy, Ordinary Installation. 


100 


100 


100 


95 


90 


85 


80 


70 


60 


50 


Probable Relative Economy. Latest Improved Installation. 


95 


98 


100 


100 


100 99 


98 


95 


90 


85 



TABLE 36. 

FLUE GAS TEMPERATURES CORRESPONDING TO FORCED CAPACITY OF BOILERS 
IN MODERN POWER PLANT INSTALLATIONS. 



Plant. 



Buffalo General Electric. . 

Cambridge Steel Co 

Commonwealth Edison Co. 

Consolidated Gas, Balti- 
more 

University of Illinois 

Detroit Edison Co 

Everett Mills 

Interborough Rapid Tran- 
sit Co., 74th St. Station. 

Narragansett Electric 
Lighting Co 

National Museum 

N. Y. Central R.R., West 
Albany 

N. Y. Central & H. R. R. R 

N. Y. Edison Waterside. . . 

Old Colony St. Ry 

Union Gas & Electric Co. . . 



Type of Boiler. 



B. & W. 
B. & W. 
B. & W. 

Edgemoor 

Locomotive 

Stirling 

Manning 

B. & W. 

B. & W. 
Geary,W. T. 

Edgemoor 
Ret. Tub. 

B. &W. 

B. &W. 

Stirling 



Rated 


Heat Sur- 


Flue 
Tempera- 
ture. 


Horse- 
power per 


face per 
Horsepower 


Unit. 


Developed. 


1140 


2.86 


705 


400 


5.14 


485 


650 


4.97 


588 


736 


4.52 


551 


328* 


2.07 


703 


2365 


4.75 


651 


130 


6.00 


599 


520 


3.00 


631 


440 


5.50 


544.2 


182 


6.40 


430 


600 


5.28 


543 


100 


4.40 


630 


650 


5.48 


550 


687 


5.25 


599 


542 


4.43 


622 



Builders' 

Rating, 

Per Cent. 



350 
194 
201 

211 
486 
211 
150 

335 

180 
155 

193 
273 
179 
190 
227 



* Assuming 10 sq. ft. of heating surface per rated horsepower. 






BOILERS 



165 



6. External Surfaces. Soot is such an excellent non-conductor of 
heat that provision must be made for its removal at frequent intervals, 
and particularly so if the boiler is expected to operate efficiently at 
heavy loads. 

These factors are treated in detail elsewhere under their respective 
headings. 

79. Effect of Capacity on Efficiency. — Tests show that if the fur- 
nace conditions are kept constant regardless of load, the efficiency of 
the boiler alone will decrease with increasing loads. But the furnace 
and grate efficiency increases with the capacity up to a certain point, 



70 

a 








































^ 


-fv*. 








350 H.P.B.&W. BOILER 

CHAIN GRATE 
OLD STYLE SETTING 
















<2;o 


'^ 
































\ 


N 






























^ 


^ 


^ 




--. 


^ 














.^^ 


py 


^ 








\ 


N 


N 


N, 




fl55 








^^ 


.^ 














^ 


\ 










y 


^ 




















k 




o 
50 




















































Jour 


w.s. 


E..F^ 


b.l7, 


1904 



U3 

^Q 

.o|g 

WW 

u u 

5 0) o 
-is D. 



0.2 



0.25 0.3 0.35 0.4 

Ptaft,over Fire in luchea of Water 



0.45 



0.5 



Fig. 62. Relation between EflSciency and Capacity. 



u 

<Si 

a 
© 

S 
s 



60- 









































































. 






A 










■ 


-»- 




- 











\ 




















































































2365 HP. STIRLING BOILER 

TAYLOR STOKER FIRED 

DELRAY STATION 





























































80 



140 160 

Per Cent of Eating 



200 



Fig. 63. Relation between Efficiency and Capacity, 



220 



166 



STEAM POWER PLANT ENGINEERING 



beyond which it remains constant or gradually drops off. For a certain 
portion of the load this increase in furnace efficiency may be at a greater 
rate than the decrease in boiler efficiency. Consequently the maximum 
combined efficiency may occur at a point either side of the rated ca- 
pacity or remain constant over a considerable range of ratings. In 



W c70 



























































































„. 


, 


.— 







_^ 














































^~- 


--- 


-^ 












































--- 


L 






















































GUARANTEED PERFORMANCE 

OF A MODERN UNDERFEED 

STOKER INSTALLATION 

Eituminons CoalrlS.OOO to 

"13,500 B.t.u. per Lb. 






























































1 1 


11 M M 













100 200 3Q0 400 

Per cent Boiler Rating 

Fig. 64. Relation between Efficiency and Capacity. 

general the combined efficiency of boiler, furnace and grate increases 
with the capacity until a maximum is reached, from which point it 
drops off steadily with each increment of increase in load. This point 
of maximum efficiency varies with the type and size of boiler, kind of 
grate, design of furnace, character of fuel and conditions of operation, 









D»_, 


.« „f 


^ 










, 1 








1 1 


"'■ 1 n 


II = 55 to 65 


"lo^ 


l"'300f« 


Ratin'g 1 




Ra 


ing 








Pro] 


Racing 
. Retort A 


reaS 


td.2 


0"Du 


tnp 13.8Sq.Ft' 




Ratios 4,7 to 69 maj 
figured fnorn fhebe 


^be 
Zlurv 


18. 














Std. 2;6'' 14.6 
Std. 2 9 15.0 K. 
XW 2 6" 19.0 3.4 
XW 29" 19.5 3.9 
XL 2'6 17.9 
XL 2'9" 18.3 
Total Area of X W - A N + K 






this ratio should be 
kept low. 














































































y 


/ 




y 


















^1^ 


,500 


?.T.X 


r. 








^ 


y 




y 




>> 

« 




^ 




. 


, 




^ 




5^ 




500 




^ 




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v'i- 




^ 










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y 




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^ 


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A 


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^ 




^ 


XW- 


=\ 


w 


1 


— 


















y 


^^ 




^^ 


^ 


^^^ 






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-ih 


■y 




^ 






























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i-^ 






























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^ 






^ 


























^ 


^ 


^ 


^ 


^ 






























^ 





































8 10 12 

H.P. per Sq.n. G.S. 



16 



18 



20 



Fig. 65. Relation between Efficiency and Capacity. (Riley Stokers.) 



BOILERS 



167 



and may range from a fraction to 200 per cent or more of the rating. 
With stokers of the underfeed type, other things being equal, the highest 
efficiency is obtained from the greatest number of retorts and the great- 
est effect on the over-all efficiency is the rate of driving per retort. 



85 



{H75 

P3 



70 



^65 



2 I 

o ~~~- 

o ^^^»^ 



2 3 4 5 6 7 

Water Evaporated per sq,.f t. of Heating 

Surface from and at 212'Talirenheit 

Fig. 66. Relation between Efficiency and Evaporation — Oil Fuel. 

The curves in Figs. 62 to 67 are based upon authentic tests and give 
some idea of the effect of capacity on efficiency in specific cases. There 
are plants throughout the country in which boilers are developing, 
during periods of peak load, capacities of 400 per cent of the rating and 
500 per cent has been realized in torpedo boat practice, but such loads 
cannot be maintained continuously with any degree of ultimate economy. 

85 



e 75 



65 






400 500 600 700 800 900 1000 1100 1200 
Boiler Horse power 

Fig. 67. Relation between Efficiency and Capacity — Oil Fuel. 



It is a question if there are thirty plants throughout the country oper- 
ating continuously day in and day out at 175 per cent rating. Widely 
varying loads are carried today in ordinary plant operation \vith over-all 
efficiencies higher than those formerly secured from constant loads and 
test conditions. 

Oil fired boilers cannot be forced to any great extent economically 
because the heat is localized and intense and severe on the boiler. This 



168 STEAM POWER PLANT ENGINEERING 

is due primarily to the fact that oil involves surface combustion while 
coal involves a volume combustion. Increase in furnace volume will 
give increased over-all efficiency at overload but the efficiency will fall 
off at normal loads. 

For influence of draft on capacity see Figs. 150 and 151. 

80. Economical Loads. — The economical rating at which a boiler 
plant should be run depends primarily upon the load to be carried by 
that individual plant and the nature of such load. The most eco- 
nomical load from a commercial standpoint is not necessarily the 
most efficient load thermally, since first cost, cost of upkeep, labor, 
cost of fuel, capacity, and the like must all be considered along with 
the thermal efficiency. The controlling factor in the cost of the plant, 
that is the number of boiler units that must be installed, regardless of 
the nature of the load is the capacity to carry the maximum peak loads. 
While each individual set of plant operating conditions must be con- 
sidered by itself the following statements * give some idea of general 
practice : 

''For a constant 24-hour load, the operating capacity, to give the 
highest over-all plant economy, is between 125 and 150 per cent of the 
boiler's normal rating. 

For the more or less constant 10- or 12-hour a day load, where the 
boilers are placed on bank at night, the point of maximum economy 
will be somewhat higher, probably between 150 and 175 per cent of 
the boiler's rated capacity. 

The third class of load is the variable 24-hour load found in central 
station work. 

Modern methods of handhng loads of this description, to give the 
best operating results under different conditions of installation, are as 
follows : 

1. The load on the plant at any time is carried by the minimum 
number of boilers that will supply the power necessary, operating these 
boilers at capacities of 150 to 200 per cent or more of their normal 
rating. Such boilers as are in service are operated continuously at 
these capacities, the variation in load being cared for by varying the 
number of boilers on the line, starting up boilers from a banked condition 
during peak load periods and banking them after such periods. This is, 
perhaps, at present the most general method of central station operation. 

2. The variation in the load on the plant is handled by varying 
the capacities at which a given number of boilers are run. At low 
plant loads the boilers are operated somewhat below their normal 

*"The Boiler of 1915," A. D. Pratt, Trans. International Eng. Congress, 1915. 



BOILERS 169 

rating, and during peak loads, at their maximum capacity. The 
ability of the modern boiler to operate over wide ranges of capacities 
without appreciable loss in efficiency has made such a method prac- 
ticable. 

3. The third method of handling the modern central station load is, 
perhaps, only practicable in large stations or groups of inter-connected 
stations. Under this method, the plant is divided into two parts. 
What may be considered the constant load of the system is carried by 
one portion of the plant, operating at its point of maximum economy. 
Due to the possibiUty of very high over-all efficiencies at high boiler 
capacities where the load is constant, where the grate and combustion 
chamber are designed for a point of maximum economy at such capac- 
ities, and where there are installed economizers and such apparatus as 
will tend to increase the efficiency, the capacity at which this portion 
of the plant is today operated will be considerably above the 150 per 
cent given as the point of highest economy for the steady 24-hour load 
for boilers without economizers. 

The variable portion of the load on a plant so operated is carried by 
the second division of the plant under either of the methods of operation 
just given." 

Standardization of Boiler Operating Conditions: Jour, A.S.M.E., Dec, 1916, p. 29. 
Operation of Large Boilers: Trans. A.S.M.E., Vol. 35, 1913, p. 313. 

81. Influence of Initial Temperature on EflBciency. — In general the 
higher the initial temperature of the furnace the greater will be the 
efficiency of the heating surface, since the heat transmitted increases 
with the difference of temperature between the water and the products 
of combustion. If the heating surface is properly distributed so that 
the final temperature of the escaping gas remains constant, the effi- 
ciency of the boiler and furnace will increase as the initial temperature 
increases, though not in direct proportion. This is on the assumption 
that the amount of heat generated per hour is the same throughout 
all ranges in temperatures. With a condition where the amount of 
heat generated remains constant and the initial temperature varies, 
the final temperature of the escaping gases remains practically constant, 
and in such cases high initial temperatures are productive of high 
boiler and furnace efficiencies. In practice these conditions are seldom 
realized and high furnace temperatures are not necessarily productive 
of high boiler and furnace efficiencies. Some tests show a decided 
gain in efficiency with the higher furnace temperatures (''Some Perform- 
ances of Boilers and Chain-grate Stokers, with Suggestions for Improve- 
ments," A. Bement, Jour. West. Soc. Engrs., February, 1904), and 



170 



STEAM POWER PLANT ENGINEERING 



others show Uttle if any improvement ('^ A Review of the United States 
Geological Survey Fuel Tests under Steam Boilers," L. P. Breckenridge, 
Jour. Wes. Soc. Engrs., June 1907). The majority of high efficiency 
records, however, are associated with high furnace temperatures since 
the latter are realized only by minimum air excess and efficient com- 
bustion. 

83. Thickness of Fire. — For a given boiler equipment, quafity and size 
of fuel and intensity of draft, a certain depth of fuel will give maxi- 
mum efficiency. Too thin a fire results in an excess of air and too 























\ 
























4UU 


















e 




bai 


)aci 


Cj- 






















s* 












/ 


y 


















'X 
















200 5A 










/ 


/ 
























V 


N, 










►. 






/ 


/ 






























X 










i " 






/ 


































N 


s , 


'— 


100 1 
















o 




iiffi 


ner 


cy 




> 


— 




-^ 




















/ 


y' 




























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3« 








/ 


/ 




































SI 






/ 


' 














































V 










































o 30 


















































2 








3 




cTl 


lick 


I 
nes£ 


iOf 


Eirt 


5, In 


Che 


s 






5 






' 


r 





Fig. 68. 



Effect of Thickness of Fire on the Capacity and Efficiency of a 350-Horse- 
power Stirling Boiler, Equipped with Chain Grate. 



thick a fire in a deficiency, the economy being lowered in either case. 
On account of the number of conditions upon which the proper thick- 
ness depends, it can only be determined for a particular case by actual 
test, the available data being insufficient for drawing conclusions. 




laiickness of Eire (Inches') 



Fig. 69. Relation between Thickness of Fire and Efficiency of Boiler Furnace and 
Grate; 512 Horsepower B. & W. Boiler and Chain Grate. 



BOILERS 



171 



The curves in Fig. 68 are plotted from a series of tests made on a 350- 
horsepower Stirling boiler equipped with chain grate at the power 
plant of the Armour Institute of Technology. The damper was left 
wide open throughout the test and the speed of the grate kept con- 
stant. Ratio of grate to heating surface, 1 to 42. Carterville washed 
coal No. 4 was used in all tests. The curves in Fig. 69 refer to the 
performance of a 512-horsepower Babcock & Wilcox boiler equipped 





~" 


































































~ 






'" 






• AAA 


















































































































































































































































- 














































































800 




























































































































































































































































































































600 




















































































































'' 






























































y 


t^ 


























"" 


^ 


■^ 


L, 








































^ 


y 












p.^ 






























'^ 






^ 
















400 












1 ( 
















"• 


^ 




































■~" 




S( 










































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L 




















































































^^ 






















































































































200 





























































































































































































































































































































































































































































































































































































































































































































































































































80 






























































































































































































































































































































60 








































' 
















4 


k 
































R 




-. 


— 


h-t 


L— 


.-, 


- 




t>_ 


-. 


- 




^ 


.- 








































































































































































































40 




































































































































A Boilec 14 Tubes High 
B - 9 .. 


— 


















































































































































_ 


_ 
























J< 


3U1 




w 


s. 


K. 


U 


^ 


» 


_ 



7 8 9 10 

Thickness oC Fire In Inches 



13 



FiCx. 70. Effect of Thickness of Fire on the Capacity and Efficiency of a 500-Horse- 
power Babcock & Wilcox Boiler. 

with chain grate and located at the power plant of the University of 
Illinois, Urbana, Illinois. The curves in Fig. 70 are plotted from a 
series of tests on a 500-horsepower Babcock & Wilcox boiler equipped 
with chain grate at the Fisk Street station of the Commonwealth 
Edison Company, Chicago, 111. In these tests the conditions of oper- 
ation are not exactly comparable, but they serve to show the variation 
of economy with thickness of fire in each case. In general, with natu- 
ral draft, fine sizes of coal necessitate thin fires, since they pack so 
closely as to greatly restrict the draft. Thin fires require closer at- 
tention to prevent holes being burned in spots, and respond less readily 
to sudden demands for steam, but have the advantage of letting the 



172 



STEAM POWER PLANT ENGINEERING 



air required pass through the grate, whereas thick fires often require 
air to be supplied above the grate to insure complete combustion. 
Thick fires require less attention and hence are preferred by firemen. 
Where sufficient draft is available thick fires are more efficient than 
thin ones, as the air excess is more readily controlled. 

TABLE 37. 

TEMPERATURE DROP OF GASES THROUGHOUT BOILER. 
(650 Hp. B. & W. Boiler, Waterside Station of N. Y. Edison Co.) 





Boiler 
and Grate 
Efficiency. 


Temperatures, Degrees Fahr. 


of 
Rating. 


Furnace 
Temper- 
ature. 


Middle 

First 
Pass. 


First 
Pass. 


Second 
Pass. 


' Middle 
Second 
Pass. 


Middle 
Third 
Pass. 


Flue. 


117.3 
126.7 


78.5 
79.6 
79.8 
77.1 
75.3 
77.6 
76.5 
72.7 


2336 


866 
893 
888 
889 
913 
956 
939 
1051 


655 
689 
681 
682 
694 
723 
700 
751 


619 
646 
633 
642 
655 
660 
634 
700 


511 
526 
521 
519 
512 
546 
523 
578 


471 

485 
481 
479 
486 
512 
493 
541 


455 
473 


128.6 
131.0 
131.0 
137.4 


2420 
2455 
2430 


468 
468 
473 
492 


142.2 




475 


185.3 


2530 


519 



83. Cost of Boilers and Settings. — The total cost of a boiler depends 
primarily on the cost of material and the cost of construction. The 
cost of material is almost a direct function of the weight but the cost 
of construction is relatively larger for small boilers than for large ones, 
so that the total cost is not a direct function of the rated horsepower. 
Furthermore, the difference in rating for various types of boilers (based 
on the extent of heating surface) has a direct influence on the cost 
per rated horsepower. For instance, Scotch-marine boilers are ordi- 
narily rated at 8 square feet of heating surface per horsepower, water 
tube boilers at 10 square feet and small fire tube boilers at 12 square 
feet. Again, the rated horsepower is independent of the working 
pressure but the latter influences the weight of material so that costs 
expressed in terms of rated capacity are widely discordant and do 
not permit of accurate formulations. The cost of a boiler ranges 
from 7.5 cents per pound for the smaller sizes to 3.5 cents for very 
large units but even this range is subject to the market price of the 
raw material. A rough rule is to allow a cost of one dollar per square 
foot of water heating surface The following equations (Boiler Room 
Economics, Patter & Simmering, Bui. 44, Kansas State Agricultural 
College) may be used as a guide in approximating the cost of different 
types of boilers with raw material based on 1915 prices. 



BOILERS 173 

Vertical fire-tube boilers; 100 pounds working pressure or less. 

Cost in dollars = 51.5 + 3.62 X rated horsepower. (43) 

Portable locomotive type fire-tube boilers; 100 pounds working pres- 
sure or less. 

Cost in dollars = 121 + 5.68 X rated horsepower. (44) 

Horizontal fire-tube boilers; 125 pounds working pressure for 100 
horsepower or less. 

Cost in dollars = 5.8 X rated horsepower — 20 for 100 to 

225 horsepower. (45) 

Cost in dollars = 211 + 3.35 X rated horsepower. (46) 

Vertical water tube boilers (100 horsepower and over); 125 pounds 
working pressure. 
Upper limit : 

Cost in dollars = 1032 -f 7.68 X rated horsepower. (47) 

Lower limit : 

Cost in dollars = 797 + 6.17 X rated horsepower. (48) 

Average : 

Cost in dollars = 912 + 6.98 X rated horsepower. (49) 

* Horizontal water tube boilers; 125 pounds working pressure. 

Cost in dollars = 149 + 8.24 X rated horsepower. (50) 

The cost of plain setting may be roughly estimated as follows: 
Horizontal water tube boilers: 

Cost in dollars = 400 + 0.8 X rated horsepower. (51) 

Return tubular boilers: 

Cost in dollars = 300 + 0.7 X rated horsepower. (52) 

For other data pertaining to cost of boiler equipment and cost of 
operating see Chapter XVIII. 

84. Selection of Type. — Boilers constructed by builders of good re- 
pute are usually designed for safety, durability, and capacity, and 
rigid specifications and inspection of material and workmanship on 
the part of the purchaser are ordinarily not necessaiy, as the makers' 
reputations are sufficient guarantee of their worth. Marked depar- 
ture from standard designs must necessarily be specified, but in most 
cases instructions are hmited to the working pressure, extent of heating 
and grate surface, the character of the furnace, and arrangement of 

* Add 10 per cent for working pressures of 200 lbs. per sq. in. Add 30 per cent 
for working pressures of 300 lbs. per sq. in. 



174 STEAM POWER PLANT ENGINEERING 

setting. Numerous tests on various types of boilers show practically 
the same efficiency provided the furnaces and boilers are properly 
designed, so that the relative merits may be considered with reference 
to (1) durabiUty; (2) accessibility for repairs; (3) faciUty for cleaning 
and inspection; (4) space requirements; (5) adaptability to the type 
of furnace and stoker desired; (6) overload capacity; and (7) cost of 
boiler and setting. For high pressures 150 pounds per square inch or 
more, the water-tube or some form of internally fired boiler in which 
the shell plates are not exposed to the high temperature of the furnace 
are considered safer than the horizontal tubular boiler because the 
shell plates and the seams of the latter must be of considerable thick- 
ness in large units, and being exposed to the hottest part of the fire 
are likely to give trouble, especially if the water contains scale or sedi- 
ment-forming elements. In the modern central station steam pres- 
sures of 200 to 250 lb. per square inch are standard practice. In a 
few recent installations pressures for 350 pounds have been specified 
and it is not unlikely but that pressures of 400, 500 and even 600 
pounds may be in immediate prospect. (See paragraph 179 for a 
discussion of high pressures.) Return tubular and stationary loco- 
motive boilers are seldom made in sizes over 250 horsepower and hence 
are not to be considered for large units. For sizes under 150 horse- 
power where overhead room is limited the return tubular boiler is most 
commonly installed, unless high pressure is essential, in which case 
the internally fired Scotch-marine boiler or special types of water 
tube boilers such as the Worthington are used. The water-tube boiler 
is usually employed in large central stations for high-pressure units of 
300 to 2500 horsepower. The particular type of water-tube boiler is 
to some extent a matter of personal taste on the part of the engineer, 
but due consideration should be given to the special requirements as 
listed above. For small powers and for intermittent operation, small 
vertical or horizontal fire-box boilers have the advantage of low first 
cost. The small air leakage and radiation losses give internally fired 
boilers an advantage over the brick-set externally-fired fire-tube or 
water-tube types, but this is partly offset by the greater extent of 
regenerative surface in the setting of the latter. In several recent 
installations the brick settings are completely encased in steel and a 
layer of high grade insulating material is placed between the brick- 
work and the casing.* This reduces the leakage and radiation losses 
to a minimum and the setting remains effective over a long period 
of time. Internally fired boilers are more expensive than the exter- 

* See "Insulation of Boiler Settings," Joseph Harrington, Power, Mar. 27, 1917, 
p. 410. 



BOILERS 175 

nally fired, though the extra cost of setting and foundation in the 
latter may bring the total cost of the entire equipment to practically 
the same figure. The design and installation of the boilers and fur- 
naces should be left at the outset to a capable engineer. 

Makers usually request the following information from intending 
purchasers : 

1. Steam pressure desired. 

2. The quantity of steam demanded. 

3. The kind of fuel to be burned. 

4. The type of furnace or stoker. 

5. The nature and intensity of draft. 

6. Nature of setting. 

7. Quality of feed water. 

85. Grates. — Grates may be divided into three general classes, 
namely, stationary, rocking, and traveling grates. The latter are 
treated in Chapter IV. Stationary grates are generally made of cast- 
iron sections in a variety of shapes as illustrated in Fig. 71. The bars 
are ordinarily from 3 inches to 4 inches deep at the center (this makes 
them strong enough to carry the load caused by the weight of the fuel 











COMMON BAR 


WMMMA 


mimmi { 


1 — """" 

• • • 


TUPPER HERRINGB 

v.v.v.v.v.-.v-v.vJ r 


ONE 


• • • 


SAW 
Fig. 71. Type 


-DUST 

s of Grate Bars. 


•• •• 

1 C^j 1 



without sagging even when the top is red hot), f inch wide at the top, 
and taper to | inch at the bottom to enable the ashes to drop clear. 
The width of the air space is determined by the size of the fuel to be 
used, the average proportions being given in Table 38. 

The "Tupper" and ''Herringbone" grate bars are stiff er and less 
likely to warp than the common form, but are not so readily sliced and 
therefore not so convenient with coal that clinkers badly. Sawdust 
or pinhole grates are used in burning sawdust, tanbark, and very 
small sizes of coal. Grates are often set horizontally and tlie bars are 



176 



STEAM POWER PLANT ENGINEERING 



held in place simply by their own weight, but long grates are best placed 
sloping toward the rear to facilitate firing. The front of the grate 
when designed for bituminous coal is often made solid, this portion 
being called the ''dead plate." It serves to hold the green fuel until 
the hydrocarbons have beejn distilled off, when the charge is pushed 
back on the open grate at the time of next firing. The length of a 
single bar or casting should not exceed three feet. The length of grate 
may be made of two or three bars and should not exceed 6 feet with 
bituminous coal, as this is the greatest length of fire that can be readily 
worked by a stoker. With buckwheat anthracite furnaces 12 feet in 
depth are not unusual, as anthracite fires require no slicing. 



TABLE 38. 
AIR SPACES AND THICKNESS OF GRATE BARS. 



Size and Kind of Coal. 



Screenings 

Anthracite — 

Average 

Buckwheat .... 

Pea or nut ..... 

Stove 

Egg 

Broken 

Lump 

Bituminous, average 
Wood — 

Slabs 

Sawdust 

Shavings 




Thickness of 
Grate Bars. 

(Inch) 
I 

f 
f 
i 
I 
I 
b 
f 
f 



The disadvantage of using stationary grates is that the fire is not 
easily cleaned. Unless the air spaces are kept free of cHnkers and 
ashes, combustion is hindered and the fire rendered sluggish. Frequent 
cleaning, however, is wasteful of fuel and reduces the furnace efficiency 
by letting in a large excess of air every time the fire door is opened. 

86. Shaking Grates. — Shaking grates have the advantage of per- 
mitting stoking without opening the fire door and require less manual 
labor than stationary grates. There is a great variety of sectional 
shaking grates on the market and some of them are made self-dumping. 
One of the best-known types is illustrated in Fig. 72. Each row or 
section of grate bars is divided into a front and a rear series by twin 
stub levers and connecting rods. An operating handle is adapted to 
manipulate either one or both of the levers in such a manner that the 



BOILERS 



177 



front and rear series may operate separately or together. The shaking 
movement causes no increase in the size of the openings and hence pre- 
vents the waste of fine fuel. Ordinarily the width of the grate is made 
equal to two or more rows of grate bars so that the live fire may be 




'^Sz 



( { { / ( { ( ( { ( ( ^ 



Fig. 72. A Typical Shaking Grate. 

shoved sidewise from one row to the other when cleaning. A depth of 
fire of from 6 to 10 inches is carried according to the nature of the 
fuel and the available draft. 

Grate Bars: Engr. U. S., Nov. 1, 1906, p. 728, Jan 1, 1907, p. 68. 

87. Blow-offs. — Boilers must be provided with blow-off pipes for 
draining off the water and for discharging sediment and scale-forming 
material. The ''bottom blow" is ordinarily 
an extra heavy pipe of suitable diameter 
connected to the lowest part of the boiler 
and fitted with a valve or cock, or both. 
The generally approved method of arrang- 
ing the blow-off pipe is shown in Fig. 89. 
This method of protecting the pipe from 
the direct action of the heated gases by 
means of a V-shaped brick pier permits 
easy examination of the blow-off through 
the cleaning door in the rear wall of the 
setting. Where boilers are arranged in 
batteries the battery may have a common 
outlet for the blow-off pipes as illustrated 

in Fig. 517. Usually the blow-off pipes may discharge into the open 
air, but this is not permissible in large cities nor is it lawful to blow 
directly into the sewer. In this case the water and sediment may be 




Fig. 



73. Blow-off Tank and 
Connections. 



178 



STEAM POWER PLANT ENGINEERING 



discharged into a blow-off tank and permitted to cool before delivery 
to the sewer, as illustrated in Fig. 73. 

Blowoff Valves and Systems: Power, July 1, 1916, p. 565. 

''Surface blows" are often installed to remove scum, grease, and 



floating or suspended particles of 



w/^/y^//yyy/y^^^^y^^^^y'^/y/yz^. 




Fig. 74. Surface Blow-off. 



dirt. The bell-mouthed shape 
shown in Fig. 74 permits the 
skimmer to accommodate itself 
to varying water level, and it is 
sometimes provided with a float 
and with a flexible joint. Fig. 75. 
88. Damper Regulators. — For 
maximum furnace efficiency the 
draft must be regulated to burn 
just enough fuel to supply the 
steam required. Where forced 
draft is employed this is done by regulating the speed of the blower. 
With natural draft it is the usual practice to regulate the draft by 
means of dampers placed in the uptake, and in order that the regula- 
tion may be effective it should be automatic. Automatic dampers 
are economical and useful and are particularly desirable in small plants 
where the demand for steam fluctuates rapidly. There are several suc- 
cessful types on the market, 
some operated by water pres- 
sure, and others by direct 
boiler pressure and in some 
of the later type by ther- 
mostats. Fig. 76 illustrates 
a typical hydraulic mecha- 
nism. Full boiler pressure 
acting at all times on a dia- 
phragm A raises or lowers a 
weight W attached to arm D 
according as the steam pres- 
sure increases or decreases. 
Arm D actuates a small valve 
y which controls the supply of water to chamber B. A diaphragm in 
chamber B raises and lowers the damper as the water pressure varies, a 
drop of 0.5 pound being sufficient to open the damper to its maximum. 
The steam diaphragm has a movement of only 0.01 inch and the water 
diaphragm 0.5 inch. When properly adjusted and given proper atten- 
tion automatic dampers work in a very satisfactory manner. 




Fig. 75. Buckeye Skimmer. 





^ 


■\ 




BOILERS 

D. 

N;; 








z:? 


steam ri 


-^ Hi 


P 






:^3 1 






i 


•— 


A 


« 


Exhaust p ^ ^ 
Water Supply ^ 


b 

its 


^ 


-w— 













179 




Fig. 76. Kitts Hydraulic Damper Regulator. 

Fig. 77 shows a section through the Tilden damper regulator, illus- 
trating the principles of the steam-actuated type. The device is con- 
nected directly to the boiler by pipe A. The 
pressure on piston B is balanced by spring C 
under normal conditions of operation. Any 
variation from the normal will cause the rod 
R to move up or down, so that the dampers are 
opened or closed in proportion to the change 
in pressure. The chamber N is separate from 
chamber M, so that steam or water cannot 
come in contact with the spring. Piston D acts 
as a guide only. In a recent design of this 
device the regulator is hydraulically actuated 
and simultaneously operates the damper and 
the stoker engine thereby automatically propor- 
tioning the air and fuel supply to the load 
requirements. 

89. Water Gauges. — The water level in a 
boiler is usually indicated either by a gauge 
glass, by try cocks, or both, connected directly 
to the boiler as in Fig. 1, or to a water column 
or combination as in Fig. 77. Each gauge- 
glass connection should be fitted with a stop 
valve which may be closed in case the tube 
breaks. In large boilers these valves, usually 
of the quick-closing type, are conveniently 
operated from the boiler-room level by means 
of a chain attached to the valve stem. Self- 
closing automatic valves have been employed, one type being illus- 
trated in Fig. 79. If the glass breaks the outrush of steam forces 
the ball against a conical seat and shuts off the supply. When a new 



Fig. 77. Tilden Steam-act- 
uated Damper Regulator. 



I 



180 



STEAM POWER PLANT ENGINEERING 



glass is inserted the ball is forced back by slowly screwing in the valve 
stem. Hinged valves mechanically operated from without are con- 
sidered more reliable than ball valves, as scale is less likely to render 
steam them inoperative. Self-closing valves 

are not allowed in modern practice. 




Water 




^^ 



Drain 

Water 

Fig. 78. Simple Water Column. 



\^ 



Fig. 79. Water Gauge with Self- 
closing Valve. 



Try Cocks 



Try cocks or gauge cocks are set at points above and below the desired 
water level, preferably connected directly to the boiler shell, but some- 
times to a water column as in Fig. 78. 
The water level is ascertained by opening 
the cocks in succession. 
NBr"^Hitll" .^rn^^^l' ^^^ objection to the latter arrangement 

^^^nj^ irrr ^^^i^ jg ^1^^^ accident to or a stoppage of the 

piping renders both gauge glass and try 
cocks useless. Water columns should be 
blown out once a day, and the gauge cocks 
opened to see that the height of the water 
indicated tallies with that shown by the 
glass. Some engineers prefer two separate 
columns to each boiler and no cocks, others 
rely solely upon cocks. 

The water column shown in Fig. 80 has 
an alarm whistle, controlled by two floats, 
which gives a high- and low-water alarm. 
Numerous devices of this class are on the 
market but they are usually regarded as 
unreliable and most engineers are content 
to depend upon water gauge and try cocks. See Power, Mar. 13, 1917, 
p. 358, for a description of a water-level indicator for high-pressure boilers. 




Fig. 80. Combined Water 
Column and High- and Low- 
water Alarm. 



BOILERS 181 

Water Gauges and Columns: Mach., Sept., 1905, p. 31; Power, Aug., 1905, p. 483; 
Am. Elecn., July, 1904, p. 359; Engr. U. S., Jan. 1, 1907, p. 58. 

90. Fusible or Safety Plugs. — Fusible or safety plugs as illus- 
trated in Fig. 81 are brass plugs provided with a fusible metal core. 
They are inserted in the shell or tubes at the lowest permissible water 
line. When covered by water the heat is conducted away sufficiently 
fast to keep the temperature below the fusing point, but when uncovered 
the low conductivity of the steam prevents the rapid withdrawal of heat, 



Inside-Types ^ ^ GVutside-Types- 




Fig. 81. Types of Fusible Plugs. 

whereupon the alloy melts and the blast of escaping steam gives warn- 
ing. The melting point of fusible metals being sometimes uncertain, 
plugs occasionally blow out without apparent cause and at other times 
fail to act when shell is overheated. Fusible plugs are required by law 
in many cities. 

91. Soot Blowers, Tube Cleaners, Etc. — Aside from the assurance 
against burning out of tubes due to the accumulation of scale, the 
maintenance of clean heating surfaces is one of the most important 
problems in connection with recent developments with higher boiler 
ratings and in the operation of large boiler units. Efficiency and 
capacity depend to a greater extent upon cleanliness (both internal 
and external) of the heating surfaces than is ordinarily reaUzed. Soot 
is an excellent heat insulating material and consequently any ap- 
preciable deposit on the heating surfaces will reduce the rate of heat 
absorption and result in high flue gas temperatures. The gain effected 
in economy and capacity by the removal of soot varies with depth, ex- 
tent and nature of the deposit and the rate of driving. No modern 
plant is operated without periodically removing this deposit. 

Surfaces exposed to the action of the products of combustion are 
customarily freed from soot and clinkers by steam lances, soot blowers 
incorporated within the setting, brushes, scrapers and similar appliances. 
Light, flocculent soot is conveniently removed at regular intervals by 
means of a hand-operated steam lance with which all surfaces are reached 
and swept clean. Under certain conditions better results are obtained 
by permanently installed soot-blowers. (See Figs. 82 and 83.) These 
consist of a series of pipes and nozzles, the latter stationary or re- 
volving, located so that all parts of the heating surface subjected to 



182 



STEAM POWER PLANT ENGINEERING 




Fig. 82. Vulcan Soot Blower Installed in Front End of a Return Tubular Boiler. 




Fig. 83, Diamond Soot Blower Applied to a Stirling Boiler, 



BOILERS 



183 



soot deposit may be swept with a jet of steam. The controlUng valves 
are located outside the setting. (See also paragraph 35.) With certain 
grades of coal under heavy furnace capacity the particles of ash and 
slag carried along with the products of combustion are in a plastic state 
and adhere to the two or three lower tubes. The accumulation may 
eventually result in a complete choking up of the gas passages. Blow- 
ing by hand lances and machine blowing devices will not remove the 
accumulation and dislodging the deposit with pokers after the furnace 
has been partially cooled appears to be the only practical solution of 
the problem. 

B&iler Cleaning Costs: Power, May 23, 1916, p. 741. 

Keeping Boiler Heating Surface Clean: Textile Wld, Sept. 9, 1916, p. 3899; Elec. 
Wld., May 20, 1916, p. 1182; Power, Aug. 31, 1915, p. 314, July 13, 1915. p. 48. 

The question of preventing the formation of scale by purification 
of the feed water and the loss in heat transmission due to scale deposit 
is treated at length in Chapter XII. In the average plant furnished 
with commercially good feed water it is customary to allow scale to 
deposit for a limited period of time and then remove it mechanically 
by tube cleaners and scrapers. The principles of construction of these 




D F L 

Fig. 84. Mechanical Tube Cleaner — Hammer Tj-pe. 

devices vary widely according to the types of boilers in which they are 
used, and depend upon the nature of the duty which they must perform. 
Mechanical tube cleaners may be conveniently divided into two classes: 

1. Those which loosen the scale by a series of rapid hammer blows, 
Fig. 84. 

2. Those which cut out the scale by a revolving tool, Fig. 85. 

The hammer device is applicable to either the water or fire-tube 
type of boiler, but the revolving cutter is applicable to the water-tube 
only. Steam, compressed air, or water under pressure may be used 
as the motive of power, though the latter is the most convenient and 
satisfactory. 

Referring to Fig. 84, the hammer head J is given a rapid motion, 
which may reach 1500 \ibrations per minute, and subjects the tube 



184 



STEAM POWER PLANT ENGINEERING 



to repeated shocks, thereby cracking the brittle scale and jarring it 
loose from the water surface of the tube. The cleaner head is at- 
tached to a flexible pipe of sufficient length to enable it to be pushed from 
one end to the other. Even if carefully manipulated the hammer is apt 
to injure the tube by swaging it to a larger diameter, producing crystal- 
lization in the metal and causing leaks where the tubes are expanded 
into thin sheets, hence its use is not to be recommended. 




Fig. 85. Mechanical Tube Cleaner — Turbine Type. 

Hydraulic turbine cutters are made in many designs, one of which is 
shown in Fig. 85. The cyhndrical casing D contains a hydrauUc tur- 
bine consisting of a fixed guide plate which directs the water at the proper 
angles upon the vanes of the turbine wheel T. The cutters C revolve 
at high speed and chip the scale into small pieces. The stream of water 
flowing from the turbine envelops the cutters, keeps their edges cool, 
and washes away the scale as fast as it is detached. Different styles 
of cutter wheels are furnished with each cleaner so as to adapt the device 
to all kinds of scale formations. In well managed plants scale is not 
permitted to deposit to a thickness greater than ^^ to yV of an inch. 

American Railway Mechanics Association; Report of the Committee on Boiler 
Washing: Ry. Review, June 19, 1915, p. 851. 



PROBLEMS. 

1. Given: Initial pressure 100 lb. gauge; barometer 29.5 in.; quality 98; feed 
water, 82 deg. fahr. Required boiler horsepower necessary to furnish a 50 horse- 
power engine with steam, engine uses 45 lb. per i.hp-hr. 

2. A 30,000 kw. steam turbine and auxiliaries require 12 lb. steam per kw-hr. 
at rated load; initial pressure 250 lb. gauge; barometer 30 in.; superheat 250 deg. 
fahr.; feed water 180 deg. fahr. Required the boiler horsepower necessary to furnish 
the turbine and auxiliaries with steam. If the boilers are operated at 250 per cent 
rating when supplying the turbine and auxiliaries, required the ratio of kilowatts 
of turbine rating to boiler rating. 

3. A boiler evaporates 90,000 lb. of water per hr. from a feed temperature of 210 
deg. fahr. to steam at 275 lb. absolute pressure and 200 deg. superheat. If the boiler 



BOILERS 185 

is being forced to 200 per cent rating when evaporating this amount of water, re- 
quired the extent of heating surface assuming that the normal rating corresponds to 
an evaporation of 3.5 lb. water from and at 212 deg. fahr. per sq. ft. of heating sur- 
face. Allowing 10 sq. ft. of heating surface per rated boiler horsepower, required 
the boiler rating. 

4. Determine the factor of evaporation for Problems 1 and 2. 

5. The following data were taken from a boiler test: 
Heating surface, 8000 sq. ft.; grate surface, 160 sq. ft.; 

Coal analysis (as fired) : Moisture 8 per cent; ash 12 per cent; B.t.u. per lb. 12,100. 

Weight per hr.: Water fed to boiler, 32,000 lb.; coal 4,000 lb.; dry refuse removed 
from ash pit, 720 lb. 

Temperatures: Flue gas, 480 deg. fahr.; feed water, 160 deg. fahr.; boiler room 
80 deg. fahr. 

Pressures: Steam pressure, 125 lb. gauge; barometer, 29.0 in., superheat, 100 deg. 
fahr. 
Required: 

a. Factor of evaporation. 

6. Boiler horsepower developed. 

c. Per cent of builders' rating developed (builders' rating = 10 sq. ft. of heating 
surface per boiler horsepower) . 

d. Evaporation per lb. of coal as fired: 

(1) Actual. 

(2) Equivalent. 

e. Evaporation per lb. of dry coal. 

(1) Actual. 

(2) Equivalent. 

/. Evaporation per lb. of combustible: 

(1) Actual. 

(2) Equivalent. 

g. Equivalent evaporation per lb. of combustible burned. 
h. Evaporation per sq. ft. of heating surface. 

(1) Actual. 

(2) Equivalent. 

i. Heat value of the combustible as fired. 

j. Heat value of the combustible as burned. 

k. Efficiency of the boiler, furnace and grate. ' 

I. Efficiency on the combustible basis. 

6, The following additional data were taken during the test outhned in Problem 
5. Flue gas analysis, per cent by volume: 
CO2 14.19; CO 1.42; O 3.54; N 80.85 
Ultimate analysis of coal as fired, per cent by weight: 
Carbon 66, hydrogen 5, nitrogen 1, oxygen 8, moisture 8, ash 12. 

a. Calculate: 

( 1 ) Complete heat balance . 

(2) Inherent losses. 

(3) Per cent of available heat utihzed. 



186 STEAM POWER PLANT ENGINEERING 

7. If the fuel (Problem 5) cost $3.25 per ton of 2000 lb., determine the cost of 
evaporating 1000 lb. of water from and at 212 deg. fahr. 

8. A test of an oil-fired furnace gave an actual evaporation of 12.77 lb. of water 
per lb. of oil with boiler and furnace efficiency of 82.8 per cent; boiler pressure 200 
lb. absolute, superheat 87 deg. fahr., feed water temperature 93 deg. fahr. Required 
the calorific value of the oil. 



CHAPTER IV 

SMOKE PREVENTION, FURNACES, STOKERS 

93. General. — Anthracite coal and other fuels low in volatile com- 
bustible matter can be burned in almost any type of furnace without 
the production of visible smoke; in fact, it is a difficult matter to pro- 
duce smoke with this class of fuel. On the other hand, bituminous and 
other ''soft" coals high in volatile matter can be burned smokelessly 
only in properly proportioned and carefully operated furnaces. 

The problem of smoke abatement is a comparatively simple one for 
large plants equipped with mechanical stokers and provided with 
ample draft, even for widely fluctuating loads, but for hand-fired plants 
it depends largely upon skillful manipulation by interested and efficient 
firemen. The order of intelligence demanded for this work is out of 
all proportion to the wages paid. In many small plants — and these 
are usually the most obstinate smoke offenders — the fireman handles 
as much as a ton of coal per hour by hand, besides caring for the feed 
pumps and water levels, keeping the boilers clean, and removing the 
ash. The boiler room is frequently poorly lighted and poorly venti- 
lated. It is, therefore, not surprising that the fireman seldom worries 
about the smoke problem. A better wage scale and more consider- 
ation for the fireman might do a great deal toward abating the smoke 
nuisance. 

Since the loss in heat due to visible smoke is usually less than one- 
half per cent, and seldom greater than one per cent of the heat value of 
the coal (see Table 39) it is a common statement among owners of 
power plants that it is cheaper to smoke than to operate without smoke. 
This is undoubtedly true in many cases where smokeless combustion 
can be secured only by admitting a considerable excess of air with a 
consequent loss in economy frequently greater than that due to incom- 
plete combustion and smoke, but if proper attention is given to the 
various factors involved practice shows that smokeless combustion 
can be effected with high boiler and furnace efficiency. 

The amount of solids discharged from a stack has no direct relation 
to visibility. A stack may appear smokeless to the eye and yet dis- 
charge considerable dust into the atmosphere. Furthermore, the sul- 
phur compounds resulting from the combustion of certain coals are 
eventually converted into sulphuric acid, and though invisible, are even 

187 



188 



STEAM POWER PLANT ENGINEERING 



TABLE 39. 

QUANTITY AND HEAT VALUE OF SOLIDS IN VISIBLE SMOKE. 
(BITUMINOUS COAL.) 

From the Report of the Chicago Association of Commerce Committee of Investigation on Smoke Abate- 
ment. (1912.) 



Test Number. 



Smoke Density, 
Per Cent. 



Solids in Visible Smoke. 



Per Cent by Weight of 
Fuel Fired. 



Per Cent of the Heat 
Value of the Fuel Fired. 



Fires with High Smoke Density. 



3 


21.97 


0.83 


0.28 


17 


20.00 


0.75 


0.36 


10 


20.00 


1.10 


0.95 


30 


15.80 


0.65 


0.49 


29 


14.50 


0.82 


0.49 


Average 


18.45 


0.63 


0.51 



Fires with Low Smoke Densitj^ 



56 
57 
80 
81 
85 
Average . 






0.51 


0.21 





0.30 


0.08 





4.07 


0.74 





1.81 


0.48 





0.47 


0.11 





0.47* 


0.32 



Average of 10 plant tests not including Test No. 



TABLE 40. 

CHEMICAL COMPOSITION OF THE SOLID CONSTITUENTS OF SMOKE. 
(CHICAGO ASSOCIATION OF COMMERCE.) 



Per Cent of Total Solids. 



Kind of Fuel. 



Hydro- Combustible 

carbons Solids 

(Tar). (Carbon). 



Mineral 

Matter 
(Ash). 



Sulphur. 



Total. 



High-pressure Plants. 



Pocahontas 
Bituminous 

Anthracite . 
Pocahontas 
Bituminous 



3.08 
4.19 



41.45 
32.80 



52.39 
59.93 



3.08 
3.08 



100 
100 



Low-pressure Plants. 



0.73 
11.43 
31.43 



31.88 
54.90 
44.06 



67.39 
33.47 
22.12 




20 
2.39 



100 
100 
100 






I 



SMOKE PREVENTION, FURNACES, STOKERS 189 

more objectioiuihlc than visible smoke from a standpoint of atmospheric 
pollution. Sulphuric acid has a disintegrating action on building material 
and produces deleterious effects upon furnishings, clothing and mer- 
chandise. Evidently smokeless combustion in itself does not prevent 
the escape of objectionable matter from the chimney, but it is a step 
in the right direction. All solid matter may be removed from the 
products of combustion by electrical precipitation * and all solids and 
gaseous sulphur compounds may be completely eliminated by ''wash- 
ing,"! ^ut these processes have not yet been developed to a basis where 
the results are compatible with the expense, at least for average power 
plant service. 

In locomotive practice from 3 to 25 per cent of the weight of dry 
coal is discharged in the form of cinders, the lower figure for a pressure 
drop of approximately 1.5 inches of water and the higher figure for a 
pressure drop of 12 inches. See Laboratory Tests of a Consolidation 
Locomotive, Bulletin No. 2, Vol. XIII, University of IlHnois, Sept. 13, 
1915, p. 25. 

In order that combustion may be smokeless and efficient, the volatile 
gases and separated free carbon must be brought into intimate contact 
with the proper quantity of air and maintained at a temperature above the 
ignitio7i point until oxidation is complete before they are brought into con- 
tact with the heat-absorbing surfaces of the boiler. Mere excess of air will 
not effect smokeless combustion, even if the gases and air are thoroughly 
mixed, if the temperature is prematurely reduced below that necessary 
for combustion by contact with the heat-absorbing surfaces of the boiler. 

Smoke may be produced, therefore, by 

1. An insufficient amount of air for the perfect combustion of the 
volatile gases. This is primarily a function of the draft. 

2. An imperfect mixture of air and combustible. 

3. A temperature too low to permit complete oxidation of the volatile 
combustible. 

Table 41 gives an idea of the distribution of smoke production by 
various industries in Chicago, in 1912. Rigid enforcement of the smoke 
ordinance has reduced the nuisance to a considerable extent, so that the 
distribution at this date differs somewhat from that shown in the table. 

The term ''smoke consumer" is a misnomer, since the combustible 
solids in visible smoke when once discharged from the furnace cannot 
be economically burned. The so-called smoke consumers are in reality 
smoke preventers. 

* Problems in Smoke, Fume and Dust Abatement: F. G. Coterell, Publication 
2307, 1914, from the Smithsonian Report for 1913. 

t Washing Smoke from Locomotives: M. D. Franey, Power, Oct. 19, 1915, p. 561. 



190 



STEAM POWER PLANT ENGINEERING 



Smoke-preventing devices may be divided into two classes: (1) those 
which may be conveniently attached to plants already in operation 
without material modification of the furnace, such as steam jets and 
other means of mixing the air and combustible gases, admission of air 
through the bridge or side wall, and mechanical draft; and (2) those 
which are an integral part of the boiler and setting, such as mechanical 
stokers, Dutch ovens, down-draft furnaces, and fire-tile combustion 
chambers incorporated with the regular setting. 

93. Hand-fired Furnaces. — Hand-fired furnaces, as a class, are most 
obstinate smoke producers. Although they can be operated efficiently 
without the production of objectionable smoke the result depends 
more upon the fireman than upon the design of the furnace. The chief 
difficulty with hand-fired furnaces Ues in the intermittent nature of 
the firing. When a fresh charge of coal is fed into the furnace an enor- 
mous volume of volatile matter is evolved. For complete combustion 
a corresponding amount of air must be supplied and intimately mixed 
with the volatile gases before contact is made with the comparitively 
cool heating surface. In the average hand-fired furnace the combustion 



TABLE 4L 

SMOKE DISTRIBUTION IN CHICAGO PLANTS. (1912.) 

(From the Report of the Chicago Association of Commerce Committee of Investigation on Smoke 

Abatement.) 



Service. 



Relative 
Weight of 
Coal and 
Coke* Con 

sumed 
During the 
Year 1912 

Tons. 



Steam locomotive 

Steam vessels 

High-pressure steam and other 
stationary plants 

Low-pressure steam and other 
stationary plants 

Gas and coke plants 

Furnace for metallurgical, manu- 
facturing and other processes. 
Total 



13.27 
0.44 

43.13 

21.91 
1.20 

20.05 



100.00 



Relative Contribution, Per Cent. 



To 
Visible 
Smoke. 



22.06 
0.74 

44.49 

3.93 
0.15 

28.63 



100.00 



To 
Solid 
Con- 
stitu- 
ents of 

the 
Smoke. 



7.47 
0.33 

19.34 

8.60 
0.00 

64.26 



100.00 



To Gas- 
eous Prod- 
ucts of 
Combus- 
tion. 



10.31 
0.60 

44.96 

23.00 
0.00 

21.13 



100.00 



To Gas- 
eous Car- 
bon Con- 
stituents 
in the 

Smoke. 



10.11 
0.55 

40.68 

23.06 
0.00 

25.60 



100.00 



To 

Sul- 
phur 
Constit- 
uents 
in the 
Smoke. 



18.22 
0.45 

53.70 

19 73 
00 

7.90 



100.00 



Total fuel consumed 21,208.886 tons. 



chamber is so small and the heating surface is so close to the grate that 
the partly burned gases strike the heating surface before oxidation is 
complete and combustion is hindered or even completely arrested. 



SMOKE PREVENTION, FURNACES, STOKERS 191 

The majority of so-called smoke preventers are merely devices for me- 
chanically mixing the air and volatile gases. These include fire-brick 
piers, baffles, arches, and steam jets. There is no question as to the 
value of these mixing devices if properly installed, but the personal 
element is too variable a factor for consistent results and the ultimate 
solution hes in mechanical stoking. The most economical and smoke- 
less hand-fired plants are those that approach the continuous feed of 
the mechanical stoker. The following rules formulated by Osborne 
Monnett, former Chief Smoke Inspector of the City of Chicago, apply 
to hand-fired furnaces burning IlUnois coal which is high in volatile 
matter. (Power, Aug. 11, 1914, p. 207.) 

Building Fires. 

Cover the grate with coal to a depth of about 4 in., and build a wood 
fire on top, or throw live coals from another boiler into the furnace at 
the bridge wall. The green coal underneath will then ignite and burn. 
As the coal in the front gradually ignites, the volatile matter must 
pass over the fire already on the grates. A live, good steaming fire 
can be built up in this way without producing dense smoke. Keep 
the doors cracked and the panels wide open, giving the fire sufficient 
air and allowing the fresh coal to be ignited from the top. This will 
keep the smoke down to a minimum. When the coal has fully ignited 
and has been well coked, fire six or eight scoops of coal on one side, 
beginning at the bridge-wall and filKng up the low spots all the way to 
the front. Do not spread the coal but allow it to lie in lumps just as it 
leaves the shovel. Before the fire gets too low on the other side, and 
after the greater part of the volatile matter has been burned off from 
the last charge, fire an equal amount of coal on the other side as before, 
always keeping the panels wide open after firing. 

When enough steam has been generated, use the jets, turning them 
on before firing and leaving them on until the bulk of the volatile matter 
has been distilled off. Always use the alternate method. For smoke- 
less operation, hand-fired boilers burning coal exclusively should not 
exceed a capacity of 150 hp. 

Cleaning Fires. 

Build up the fire on one side and let the other side burn down. Just 
before cleaning, ''wing" over the live coal from the burned-out side 
and pull out the cHnker and ash, cleaning the grates thoroughly. Cover 
the clean grates with green coal and push over live coal from the other 
side. When the cleaned side has become thoroughly ignited and the 
volatile matter has passed off, throw in coal to fill the spots not covered 
and pull the clinker and ash out of the other side. Cover the grates 
with green coal as before, winging over live coals from the opposite 
side. Keep the panels wide open and allow the fresh coal to ignite 
thoroughly. Never allow the fire to burn low before cleaning if carry- 
ing a heavy load, as there is a possibiUty of losing the steam pressure. 



192 STEAM POWER PLANT ENGINEERING 

After cleaning, follow up closely with the alternate method of firing 
until the fuel bed is thick enough to hold the pressure. 

Carry as thick a fire a^ the draft will permit and do not spread the 
coal over the grates. 

Banking Fires. 

Throw 15 or 20 scoops of coal on each side. Open the panel doors 
sHghtly, close the ashpit doors and partially close the damper. To 
break up the bank, level the fire, with the panel doors open, and start 
firing by the usual method, making sure the damper is -svide open. 

Keep the fire brick by the alternate method, using the panel doors 
and steam jets, and regulate the steam pressure with the ashpit doors. 
This insures a temperature in the furnace high enough to maintain the 
brickwork at the igniting point of the coal and promotes combustion 
of the volatile matter. At the same time, it keeps down the distillation 
of the volatile matter to a low rate, and by having the damper open 
and the panels cracked, the circulation of the gases is not retarded. 
Do not try to regulate the steam pressure by the damper or smoke will 
be produced. 

The method just described is contrary to the general rules for firing, 
but its philosophy is explained in the following: If a shovelful of coal 
is thrown on a bright fire by the spreading method, every particle of 
that coal is immediately subjected to the intense heat of the fire and the 
volatile matter is rapidly driven off. If this is followed by more coal, 
the result is a volume of volatile matter which is beyond the capacity 
of the furnace to handle without dense smoke. 

If, on the other hand, the fuel is fired in a lump from the shovel with- 
out spreading, there is a considerable quantity of the coal which does 
not immediately become subjected to the high temperature. The 
coal on the outside of the pile gives up its volatile and the coal within 
is not affected until the volatile matter has been distilled from the outside 
lumps. Furthermore, the volatile matter from the inner portions of 
the pile must pass outward through the incandescent outer layer of 
fuel much in the same way as in the underfeed stoker. In this way 
the production of smoke can be retarded, and the more coal thrown on 
the fire at once, the less smoke. Two shovelfuls of coal fired by the 
spreading method on a clean, bright fire will make more smoke than ten 
shovelfuls of coal fired by the lump method. In practice, it will be 
necessary to determine just how much coal should be fired at once, 
but six or eight shovelfuls on a side with the draft operated according 
to the method ordinarily used will be found to be about right. 

When a battery of boilers is to be fired, the fires should be fed by the 
alternate method, as before, passing from one boiler to another until 
they are all charged and then repeating on the other side of the furnaces. 
It should be determined by experiment, however, just how many fur- 
naces can ])e fired consecutively without producing dense smoke, and 
after this has once been made known, the fireman should adhere strictly 
to the rules laid down in this regard. 

94. Dutch Ovens. — One of the earliest attempts at hand-fired smoke- 
less-furnace construction consisted in placing a full-extension Dutch 



SMOKE PREVENTION, FURNACES, STOKERS 



193 



Oven (Fig. 86) in front of the boiler. This provided a large com- 
bustion chamber but the setting was extravagant in floor space and the 
intense radiation from the incandescent furnace Uning effected a too 
rapid distillation of the volatile matter from the green fuel. As a result 




Fig. 86. Plain Dutch Oven — Full Extension. 

the velocity of the gases was too high to permit complete oxidation 
within the furnace and visible smoke could not be prevented from form- 
ing except at light loads. Steam jets placed at the sides of the setting 
and blowing across the fire assisted in mixing the gaseous products 




I 



LOMGITUDINAL SECTION " SECTION A-A 

Fig. 87. Dutch Oven for Burning Wood Refuse. 

but did not satisfactorily solve the problem. By placing the oven 
partly (semi-extension) or completely (flush front) underneath the boiler 
proper the extra space requirements were reduced or completely elimi- 
nated but a considerable portion of the heating surface was insulated 
from the fire at the expense of capacity. The next step was to remove 



194 STEAM POWER PLANT ENGINEERING 

part of the oven roof and expose the boiler surface to the direct action 
of the fire. This increased the economy and capacity of the setting 
but still failed to effect the desired result. The introduction of a de- 
flector arch at the end of the oven made smokeless combustion possible, 
but the setting was rather high and expensive to install and maintain. 
The final development consisted in arranging the deflector arches and 
a double-arch bridge wall as illustrated in Fig. 89. Dutch ovens are 
generally used in burning wood refuse and similar fuels. See Fig. 87. 

95. Twin-flre Furnace. — This arrangement, illustrated in Fig. 88 in 
connection with a hand-fired return tubular boiler, is a double furnace 
formed by longitudinal arches extending between bridge wall and fire 
door. 

The furnaces are fed and manipulated alternately, the object being 
to have one furnace in a highly incandescent state, while green fuel is 
fed into the other. Air is admitted both below and above the grate, 
and the volatile gases are supplied with the necessary oxygen for com- 
bustion before they come into contact with the comparatively cool 
boiler surface. 

The gases from both furnaces first pass into a chamber formed by a 
single arch sprung across the entire inner setting from the side wall, a 
short retarding arch being placed between this intermediate chamber 
and the rear of the setting. A special tile of high-grade refractory clay 
is used, the thickness varying from 4 to 6 inches, depending upon the 
size of furnace and the length of span. The furnace can readily be 
substituted for the ordinary types in common use under any standard 
tubular or water-tube boiler and may be installed either under the boiler, 
as indicated in the illustration, or in an extension Dutch oven. This is an 
excellent furnace, and when properly manipulated gives smokeless and 
efficient combustion. 

96. Chicago Settings for Hand-flred Return TubuIaF Boilers. — Figs.' 
89 and 90 show details of settings for return tubular boilers as recom- 
mended by the Chicago Department of Smoke Inspection, and which 
may be considered the latest development in hand-fired smokeless 
settings for IlHnois bituminous coals. The setting illustrated in Fig. 
89, and known as the Double-arch Bridge-wall Furnace, is intended 
for low pressure work where steam jets are not effective and where 
the rate of combustion is 15 pounds of coal per square foot of grate 
surface per hour or less, and that shown in Fig. 90 where the rate of 
combustion is greater or where the plant has a regular power load. 
The dimensions refer to a specific set of conditions and are not general. 
Both settings require careful manipulation for smokeless combustion as 
is the case with hand-fired furnaces in general. It has been the ex- 



SMOKE PREVENTION, FURNACES, STOKERS 



195 




196 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



197 




o 



198 



STEAM POWER PLANT ENGINEERING 



perience of the Department that most violations of the smoke ordinance 
are due primarily to insufficient draft, the required rate of combustion 
being too high for the available air supply. The following note refers to 
the No. 8 furnace: First grade fire brick to be used throughout with 
the exception of the combustion chamber floor and the side walls of 
the combustion chamber back from a point one foot behind the rear 
face of the wing-walls. Wing-walls to be bonded into the side walls. 
No air space to be left in the setting walls. Fire doors must provide for 
special air admission of an area equal to four square inches for each 
square foot of grate surface. 

97. Burke's Smokeless Furnace. — Fig. 91 shows sections through 
a Burke smokeless furnace as installed in a number of tall office buildings 
in Chicago. It amounts virtually to a Dutch oven equipped with 
shaking grates, and embodies an extension self-feeding coking oven of 
cast-iron section lined with fire brick and protected from overheating 
by air circulation through the sections. Natural draft is used, the 




Fig. 91. Burke's Smokeless Furnace. 



fire doors being closed; but air is admitted above as well as below the 
fire. As this stoker is manipulated by hand, more or less attention is 
required of the operator in keeping the fire clean. Furnaces of this 
type at the power plant of the Majestic Theater building, Chicago, 111., 
are giving good results. 

98. Down-draft Furnaces. — Fig. 92 shows the application of a 
Hawley down-draft furnace to a Heine water-tube boiler. In this fur- 
nace there are two separate grates, one above the other, the upper 
one being formed of parallel water tubes connected with the water 
space of the boiler through the steel headers or drums, A and D, in 
such a manner as to insure a positive circulation. Fuel is supplied to 
the upper grate, the lower one, formed of common bars, being fed by 
the half-consumed fuel falling from the upper grate. Air for com- 
bustion enters the upper fire door, which is kept open, and passes first 
through the bed of green fuel on the upper grate and then over the in- 



SMOKE PREVENTION, FURNACES, STOKERS 



199 




200 STEAM POWER PLANT ENGINEERING 

candescent fuel on the lower grate. A strong draft is required, due 
to the relatively small upper grate area and the correspondingly high 
rate of combustion. Lump coal gives better results than the smaller 
sizes, as the latter are apt to fall through the upper grate before being 
even partially consumed and when such is the case efficient results cannot 
be obtained. If carefully manipulated this furnace with fire-tiled tubes 
as illustrated in Fig. 92 gives high boiler efficiency and smokeless com- 
bustion, but its overload capacity is limited. Without the fire tiling 
smokeless combustion is possible only at light loads. 

The down-draft furnace is remarkably successful on low rates of 
combustion, 10 lb. per sq. ft. per hour or less and is used extensively 
for heating loads. 

99. Steam Jets. — The main purpose of a" steam jet in connection 
with ''smokeless furnaces" is to mix the air and gases and insure inti- 
mate mixture of the products of combustion. This action is purely 
mechanical, the steam in itself not being a supporter of combustion. 
The claims sometimes made that steam increases the calorific value of 
fuel are erroneous. There are conditions with certain grades of coals 
and refuse under which a moderate amount of steam injected into 
the furnace promotes complete combustion and increases the efficiency 
of the boiler. Such results, however, are due to increase in available 
heat and not to increase in actual calorific value. A theory advanced 
in this connection is that "hydrogen and CO formed by the reaction 
between the steam and incandescent carbon unite with the oxygen 
of the air passing through the grate and generate intense heat. This 
heat dissociates a part of the steam into hydrogen and oxygen. The 
hydrogen immediately recombines with oxygen of the air, while the 
oxygen in its nascent state effects complete combustion of the hydro- 
carbons which under ordinary conditions escape in the form of smoke. 
Although it takes as much heat to dissociate steam into its elements as 
is given off when the hydrogen burns back again to water vapor the gain 
in available heat effected by the steam hes in the combustion of the 
hydrocarbons which would otherwise be discharged up the stack. 
The heat necessary to superheat the steam to stack temperature must 
be charged against the coal pile but the loss may be more than offset 
by this increase in available heat. It takes the same amount of oxygen 
to burn the hydrogen as is liberated by dissociation so there is no extra 
oxygen available for combustion, but the oxygen thus liberated is in a 
nascent state and combines much more readily with the hydrocarbons 
than does atmospheric oxygen." 

There is no question as to the value of properly installed steam jets 
in maintaining smokeless combustion in internally fired furnaces, hand- 



SMOKE PREVENTION, FURNACES, STOKERS 201 

fired return-tubular boilers and improperly designed furnaces, but 
taking all things into consideration better results may be had with 
properly designed furnaces equipped with mechanical stokers. A 
plain setting and steam-jet equipment either manually operated or 
automatic will usually average from 8 to 12 per cent smoke density (see 
paragraph 105). The ''Chicago No. 8 Furnace" properly manipulated 
can be operated with about 2 per cent smoke density. A smokeless 
stack with hand firing is not a true indication of efficient operation, 
since the air dilution may be excessive and the heat demands of the steam 
jets may be very great. Since air requirements are greatest at the 
moment of firing fresh coal, and the demand diminishes as distillation 
of the volatile matter progresses, steam jets need close regulation for 
best economy. If permitted to run continuously, as is often the case, 
they may use considerably more of the energy of the coal than they 
save by effecting smokeless combustion. Practically all of the so-called 
''smoke consumers" for hand-fired furnaces depend upon the steam 
jet or admission of air only above the fire for their operation. In most 
of these the jets are automatic and operate independently of the fireman. 
The most efficient jets are those based on the injector or siphon prin- 
ciple in which the jet induces a flow of air along with the steam. The 
steam nozzles are usually placed in the front wall and are charged down- 
ward toward the bridge wall, as illustrated in Fig. 90. Occasionally 
they are placed in the side wall or even in the bridge wall, but the 
front wall construction appears to be the best. The majority of the 
patented smokeless furnaces involving the use of the steam jet do not 
conform with the requirements of the Chicago Department of Smoke 
Inspection, chiefly because of faulty furnace design. 

Theory, Practice and Design of Hand-fired Furnaces and Modern Methods of Smoke 
Prevention: National Engr., Nov., 1913, p. 670 (Serial). 

For valuable data pertaining to smokeless combustion including brick settings for 
all types of hand-fired and stoker-fired furnaces see serial article by O. Monnett, 
Chief Smoke Inspector, Chicago, 111., Power, May 12, 1914 to Jan. 5, 1915. 

100. Mechanical Stokers. — Uniform evolution of the volatile gases of 
the fuel is the essential requisite for smokeless combustion, and it is 
for this reason that mechanical stokers as a class are more effective in 
preventing smoke than any apparatus accompanied by intermittent 
firing. Stokers which feed irregularly have the effect of hand-fired 
furnaces, and it is necessary not only to employ some powerful auxiliary 
mixing device but also to furnish at times an extra supply of air to 
take care of the enormous volume of volatile gas evolved after a fresh 
charge of fuel is added. 

Carefully adjusted automatic stokers owe their high efficiency to: 



I 



202 STEAM POWER PLANT ENGINEERING 

(1) uniformity of feed; (2) proper proportion of air and combustible; 
(3) absence of excessive air dilution, as when the fire doors are opened in 
connection with hand firing; and (4) self-cleaning grates. 

Daily records are essential with any type of stoker or hand firing if 
efficient results are expected, as only by frequent observation is it 
possible to determine the proper adjustment of air supply, depth of fire, 
rate of feed, and the like. Control of air supply is almost as important 
as the upkeep and effective operation. In the best firing practice the 
right amount of air, depth of fire, and rate of feed must be worked out 
by the engineer. 

Stokers are often condemned by owners as inefficient and inferior to 
hand stoking because no particular attention has been paid to them 
beyond filling the hopper with coal. They should be operated in strict 
accordance with the principles of design. 

In plants of 2000 horsepower or over, the installation of mechanical 
stokers and coal conveyors effects a considerable saving of labor and 
can usually be relied upon to solve the smoke problem if reasonable 
attention is given to their operation. In smaller plants interest on the 
investment and other considerations may make hand firing more 
economical, although many stoker-fired plants of capacities as small as 
200 horsepower are giving satisfaction, particularly in places where a 
poor grade of fuel is used and smoke ordinances are rigidly enforced. 
A stoker of the self-cleaning, slow-running type requires much less 
attention than the hand-fired furnace. With hand firing one fireman 
can efficiently attend to the water, coal, and ashes of about 200 horse- 
power, or handle coal for, say, 500 horsepower, whereas with good 
automatic stokers equipped with overhead bunkers and down spouts he 
can readily take care of 4000 horsepower. 

The best stokers are those which are least compHcated and simplest 
in operation. A cheap stoker is a poor investment, since the cost of 
repairs and shutdowns will usually amount to more than the saving in 
price. 

The following outline gives a classification of a few of the best-known 
American mechanical stokers:' 





Chain Grates. 




Babcock & Wilcox 


McKenzie 


Leclede-Christy 


Green 


IlHnois 
Overfeed. 


Westinghouse 


Step Grates — Front feed. 




Step Grates — Sidefeed. 


Roney 




Murphy 


Wilkinson 




Detroit 


Acme 




Model 



SMOKE PREVENTION, FURNACES, STOKERS 



203 



Jones 
American 



Underfeed. 
Taylor 
Riley 



Down Draft. 
Hawley 



Westinghouse 
Type^'E" Combus- 
tion Engineering Co. 

Sprinkler 

Swift 

Vulcan 



Powdered Fuel. 
See paragraphs 41-46. 



101. Chain Grates. — The standard type of chain grate is one of 
the most popular forms of automatic stokers for burning small sizes of 
high ash and free-burning bituminous coals. (30 to 40 per cent volatile 
matter and 10 to 20 per cent ash.) It is also adapted to Ugnites and 
the very high ash coals of the West. With low ash coals at high rates 
of combustion the grate is apt to become overheated, and breakage with 
attendant high maintenance may result. The standard chain" grate 
embodies a moving endless chain of grate bars mounted on a frame with 
provision for the continuous and uniform feeding of coal into the furnace, 
the fuel and the grate moving together. As usually installed the surface 
of the grate is horizontal though in some designs it is given a slight 
inchne toward the bridge wall. The operations of feeding the coal, 
carrying it through the progressive stages of combustion, removing the 
ashes and clinkers, and maintaining a clean grate and free air supply 
are automatic. The driving mechanism consists of a gear train actuated 
by ratchet and pawls, the arms carrying the latter being given a re- 
ciprocating motion by an eccentric mounted on a Hne shaft. The latter 
may be driven by any type of engine or motor and the speed of the 
grate regulated by varying the stroke of the arm carrying the pawls. 
Fuel is fed into a hopper placed at the front end of the furnace and the 
depth of the fuel regulated by a guillotine damper. The front part 
of the furnace is provided with a flat or sUghtly inclined ignition arch 
the function of which is obvious. The entire grate and driving mecha- 
nism is mounted on a permanent truck and may readily be removed from 
beneath the boiler. The thickness of the fire and the speed of the grate 
should be so regulated that when the fuel has reached the end of the 
grate it shall have been completely consumed and incandescent ashes 
only will be discharged into the pit. With chain-grate stokers there 
may be considerable leakage of air between the grate and bridge wall, 
through the coal in the hoppers, under the coal-gate and through the 
fire bed at the rear where it is mostly ashes unless care is used in regu- 



204 



STEAM POWER PLANT ENGINEERING 



lating the depth of fire and ash bed and provision is made for preventing 
this ''short circuiting" of the air supply. 

In the ''IlHnois" chain grate the hve grate area may be varied by 
means of dampers placed immediately below the upper chain surface. 
This permits the use of a ''short fire" at hght loads without excessive 
air leakage. Chain grates as a class are seldom operated at loads ex- 
ceeding 250 per cent of the rated boiler capacity. 

Fig. 93 shows the general apphcation of a B. & W. chain grate to a 
B. & W. boiler. The ignition arch is parallel to the grate and covers 




/75^ 



imm 



Fig. 93. Babcock and Wilcox Boiler, Chain Grate, Ordinary Setting. 

a considerable portion of the grate surface. The bridge wall is fitted 
with a water back as indicated, to prevent the grate bars from being 
burned. With normal uniform loads this style of ignition arch and 
setting insures practically smokeless combustion, but careful manipu- 
lation is necessary with rapidly fluctuating loads to prevent the for- 
mation of objectionable smoke. 

Fig. 94 shows an application of a Babcock & Wilcox chain grate to 
a horizontal water-tube boiler as installed at the Quarry Street Station 
of the Commonwealth Edison Company. It will be noted that the 
stoker is applied to the rear of the setting. This arrangement of stoker 
and Sewall baffling effects smokeless combustion, but the life of the fur- 
nace is short because of the low spring of the arch. 

Fig. 95 shows another arrangement of a Babcock & Wilcox boiler 
and chain grate with vertical baffling as instaUed in units 5 and 6 of the 



SMOKE PREVENTION, FURNACES, STOKERS 



205 




206 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 207 

Quarry Street Station and units 1 and 2 of the Northwest Station. 
This is a smokeless setting up to 175 per cent of rating. 

Fig. 96 gives the general details of a Green Type ''K" chain grate 
as applied to a horizontally baffled water-tube boiler. This setting 
conforms with the requirements of the Chicago Smoke Department 
and is smokeless up to 200 per cent rating. 

The standard type of chain grate is not adapted to coking coals on 
account of the swelling and fusing action of the fuel under the ignition 
arch. The chain grate may be modified to burn this class of fuel by 
introducing incUned coking plates immediately under the front of the 
ignition arch and agitating them mechanically during the period of 
distillation. This agitation prevents the coal from fusing together and 
by the time the fuel reaches the grate proper it no longer tends to cake. 
The Green Type ''L" chain grate is an example of this modification. 

Chain Grates and Smokeless Settings: Power, Oct. 13, 1914, p. 532; Oct. 20, 1914, 
p. 560; Nov. 3, 1914, p. 658. 

102. Overfeed Step Grates. — In stokers of the overfeed step grate 
type coal is pushed in at the top of the slope and caked by the aid of 
a fire-brick arch and fed downward progressively by the movement of 
the grate bars aided by gravity. The upper portion of the bars is 
arranged to retain the uncaked part of the coal, changing to larger open- 
ings in the lower portion where the coal has fused and combustion is 
chiefly that of fixed carbon. The clinker collects at the bottom where 
it is crushed by rolls or dumped. Since the fixed carbon combustion 
occuj-s directly on the grate all overfeed stokers are subject to over- 
heating and destruction of the grate bars may be considerable, particu- 
larly with high sulphur coals. Any of these stokers will operate effi- 
ciently with little trouble from clinker and burning if installed in properly 
designed furnaces and operated at their proper capacity. Very few 
stokers of this type are operated at loads exceeding 200 per cent of 
the rated boiler capacity and for this reason are not much in evidence 
in the modern large central station. The Roney stoker, Fig. 97, and 
the Wilkinson stoker. Fig. 98, are examples of the frontfeed type of 
step grates. The Roney stoker consists of a hopper for receiving the 
coal, a set of rocking stepped grates inclined at a proper angle from the 
horizontal, and a dumping grate at the bottom of the incline for receiving 
and discharging the ash and clinkers. The dumping grate is divided 
into several sections for convenience in handling. The coal is fed onto 
the inclined grate from the hopper by a reciprocating "pusher" actuated 
by the ''agitator." The power is supplied through an eccentric oper- 
ated by a small engine or motor. The normal feed is about 10 strokes 
per minute. The grate bars rock through an arc of 30 degrees, assuming 



208 



STEAM POWER PLANT ENGINEERING 



2n is»«ii TO 










O 



SMOKE PREVENTION, FURNACES, STOKERS 209 







DataUs of Coostructioo of the Roner Mechanical Stoker 

Fig. 97. Details of Roney Stoker. 



THK MECHANISM OF THE WILKINSON STOKtR. 




Fk;. U.S. Details of Wilkiiisoii Stoker. 



210 STEAM POWER PLANT ENGINEERING 

alternately horizontal and inclined positions. The construction per- 
mits abundance of air to pass through the fuel, with little or no possibility 
of coal dropping through the grate. A coking arch of fire brick is sprung 
across the furnace as indicated. This stoker operates with natural or 
forced draft and, with suitable headroom, effects complete and efficient 
combustion. 

In the Wilkinson stoker the inclined grate bars are hollow and are 
arranged side by side, every alternate bar being movable. When in 
operation there is a constant sawing action of the grate bars, causing 
the fuel to flow forward and downward. A small steam jet with about 
xV-iiich opening is introduced into the end of each hollow grate bar, 
and induces the required amount of air for combustion through air 
openings approximately { inch wide by 3 inches long. These stokers are 
driven by two small hydraulic motors. The water is furnished by a 
small pump and is used over and over again. 

Front Feed Stokers and Smokeless Settings: Power, Nov. 17, 1914, p. 712. 

The sidefeed step-grate stoker is represented by the ''Murphy," 
*' Detroit," and ''Model." Fig. 99 ^ shows a front vertical section 
through a Murphy automatic stoker and furnace. The apparatus is 
in effect a Dutch oven equipped with an automatic feeding and stoking 
device. Coal is introduced either mechanically or by hand into the 
magazine at each side of the furnace and above the grate and descends 
by gravity upon the coking plate. Reciprocating stoker boxes push the 
coal upon the grate bars. Every alternate grate bar is movable and 
pivoted at its upper end. A rocker bar driven by a small motor or 
engine causes the lower ends to move up and down, this action pro- 
ducing the required stoking effect. A device for grinding up the chnker 
and ash is provided as shown at the bottom of the furnace. This is 
hollow and is connected by a 2-inch pipe with the smoke flue, so that 
the cold air passing through prevents it from being destroyed by the heat. 
Air is supplied to the green coal through flues passing under the coking 
plates, and the speed of the stoker boxes and grate bars can be regu- 
lated to conform to any rate of combustion. On account of the large 
fire-brick combustion chamber, this stoker with careful manipulation 
is capable of practically smokeless combustion. 

Side Feed Stokers and Smokeless Settings: Power, Nov. 8, 1914, p. 802. 

103. Underfeed Stokers. — This type of stoker has practically sup- 
planted all other types in the modern large central station burning 
coking bituminous coals and is adapted to all grades and sizes of free 
burning bituminous coal. The underfeeds are essentially forced draft 
stokers, since they operate with restricted air openings and very deep 



SMOKE PREVENTION, FURNACES, STOKERS 



211 




SIDE SECTION 

Fig. 99. Murphy Furnace. 



212 



STEAM POWER PLANT ENGINEERING 



fires. The forcing capacity is tremendous, reaching as high as 450 
per cent of rating. With this class of stoker headroom is the principal 
factor. For smokeless combustion special brickwork is not necessary 
and coking arches may be dispensed with entirely. Some of the best- 
known underfeeds are the Jones, Taylor, Riley, Westinghouse, Com- 
bustion Engineering Company's Type ''E, " and American. 

Fig. 100 shows the general principles of the Jones underfeed stoker. 
It consists of a steam-actuated ram with a fuel hopper outside of the 
furnace proper and a fuel magazine and auxiliary ram within. Air for 




Fig. 100. Jones Underfeed Stoker. 

combustion is admitted through openings in the tuyere blocks on either 
side of the retort. Coal is fed into hoppers and forced under the bed 
of fuel in the stoker retort, where it is subjected to a coking action.' 
After liberation of the volatile gases the coke is pushed toward the top 
of the fire. The top of the fire, nearest the boiler, is always incandescent. 
Each charge of coal is given an upward and backward movement. 
Air is admitted through the tuyere blocks at the point of distillation 
of the gases. Grate bars form no part of the Jones system, and it is 
therefore impossible for the fuel to fall through. There is no ash pit. 
The non-combustible matter is removed from the furnace by hand. 
The standard size of the retort is about 6 feet in length, 28 inches in 
width, and 18 inches in depth, and experience has shown that other 
sizes are not necessary since the spaces between retort and side wall 
of the various furnaces may be provided for by extending the width 
of the dead plates. One or more stokers are installed in each furnace, 
depending upon the capacity of the boiler and the width of the furnace. 
The steam pressure automatically controls air and fuel supply, propor- 
tioning them to each other and to varying loads in the correct degree. 
The result is that the stoker effects complete and smokeless combustion. 
The only variable element in the operation of this stoker, once it is cor- 
rectly installed, is cleaning of fires, but if the fireman is careful to burn 
down the coals before breaking them up the production of smoke may 
be avoided. Jones underfeed stokers are adaptable to all grades and 
sizes of bituminous coal, and on account of forced draft are capable of 
burning very low grades of coal. 



SMOKE PREVENTION, FURNACES, STOKERS 



213 




214 



STEAM POWER PLANT ENGINEERING 



Fig. 101 shows the general details of a Taylor underfeed stoker for 
burning bituminous coals. The device consists essentially of a series 
of alternate retorts and tuyere boxes inclined as indicated. Each 
retort is fitted with two rams — the upper for pushing the green fuel 
outward and upward and the lower one for forcing the fuel bed and 
refuse toward the dump plates at the rear. Air is supplied by a volume 
blower and enters the furnace through openings in the tuyere boxes. 
The dump plates are hung on the rear of the wind box and are con- 
trolled from the front of the stoker. Extension grates are inserted 
between the mouth of the retort and the dump plates, when the nature 
of the fuel makes this arrangement desirable. This extension may be 
rocked if necessary. In the later designs of this stoker the dump 
plates are actuated by a steam cylinder. The valve mechanism is 



Kecii)rocatine 

P ; I ', I Pu.,b^i Nose 

' for Dumping 



ifety Shearing 
Pin through 
Connecting Rod 




Shaft for Adjusting Combusiiofr 
Position of Ash Plates 

Fig. 102. General Assembly of Riley Underfeed Stoker. 

placed at the side door so that the operator can manipulate the dump 
plates in full view of the ash and clinker. The plate is so designed that 
it can be rocked without dumping, hence a similar motion in the exten- 
sion grates is unnecessary. The stoker and blower are operated by the 
same engine, the air and coal supply being automatically controlled 
by the variation in steam pressure. Taylor stokers may be operated 
smokelessly and efficiently at very heavy overloads and are much in 
evidence in the eastern states. The steam required to operate the blower 
and stoker varies from 2.5 to 5 per cent of the steam generated, depending 
upon the size of the installation and the percentage of rating developed. 
The Riley, Fig. 102, is a multiple retort stoker with an incline of 
about 20 degrees. The distinctive feature of this stoker is that the sides 



SMOKE PREVENTION, FURNACES, STOKERS 215 




Fig. 103. Duplex Furnace with Riley Underfeed Stokers. 



of the retort reciprocate relative to the bottom. This causes the fuel to 
be moved at a uniform rate out of the retort and across the ash-sup- 
porting plates until it is discharged through the adjustable openings 
next to the bridge wall. No special shape of wind box is required with 
this stoker since the air chamber is formed by the boiler side walls and 
any convenient floor The air supply may be controlled by hand or 
automatically. One man can operate ten or twelve stokers; sif tings 
are negligible amounting to 0.2 or less and the wind box and retorts 
need be cleared but once a month. The power required to operate the 
stoker is approximately J horsepower per retort. Fig. 103 illustrates 



216 



STEAM POWER PLANT ENGINEERING 



one of the latest installations of the Riley stoker for high efficiency and 
extremely high overload capacity. 

Maintenance Costs of Two 2365-hp. Boiler Units with Taylor Stokers: Trans. 
A.S.M.E., Vol. 35, 1913, p. 327. Installation Data for Underfeed Stokers: Elec. Wld., 
Nov. 18, 1916, p. 1009. 

Underfeed Stokers: Power, Dec. 15, 1914, p. 838; Jan. 26, 1915, p. 132. 

104. Sprinkling Stokers. — In this system of stoking the fuel in 
finely divided form is distributed by sprinkling uniformly over the en- 
tire area of the grate. With 
the proper adjustment of air 
supply and feed the volatile 
gases are distilled off continu- 
ously before the grate is cov- 
ered by the new coal and 
without materially lowering 
the temperature of the incan- 
descent fuel. Mechanically 
the operation involves con- 
siderable difficulty. Sprinkling 
stokers do not conform to 
Chicago requirements. 

Fig. 104 gives the general 
details of the Swift stoker, 
illustrating a commercially 
successful stoker of this type. 
The apparatus is self-contained 
and is bolted to a frame cast- 
ing in front of the setting, and 
takes the place of the fire 
door. It may be swung back 
from the fire-door opening in 
much the same manner as the 
ordinary fire door. Coal of 
nut size or smaller is fed into 







Fig. 104. 



Swift Sprinkling Stoker. 



a small hopper, of about 300 pounds' capacity, from which it gravitates 
on to a berm plate and pusher plate. By means of the latter the fuel 
is fed to rapidly revolving spreaders, which crush it into small particles 
and throw it onto the grate. The fine or powdered coal is burned in 
suspension and the heavier coal falls to the grate. The spreaders are 
heavy pieces of cast steel, revolving about a common axis and shaped 
hehcally so as to throw the fuel in a direction at right angles to the 
face of the machine. There are several of these spreaders so arranged 



SMOKE PREVENTION, FURNACES, STOKERS 217 

on the shaft that adjacent spreaders throw the fuel in different directions. 
This stoker is not self-cleansing, that is, the ashes must be removed by 
hand or by suitable shaking grates. 

105. Smoke Determination. — Smoke measurements may be either 
quantitative or relative. 

The most satisfactory method, at this writing, of determining the 
quantity of smoke passing through a chimney is that adopted by the 
Chicago Association of Commerce. A continuous sample of chimney 
gas is drawn from the stack by means of a special Pitot tube and ex- 
hauster, and the soUd particles are entrapped in a filter. The tube is 
so arranged that the rate of flow through the apparatus is the same as 
that in the chimney. Since the area of the tube opening bears a fixed 
ratio to that of the chimney, the weight of carbon, cinders, soot and the 
like caught in the tube filter is a measure of the total weight emitted 
from the stack. 

Quantitative measurements are of considerable value in estimating 
the amount of energy lost in the production of visible smoke, but are 
seldom attempted in regular practice. 

There are several methods of determining smoke, relatively. The 
most common is that devised by Ringelmann, and is commercially 
known as the Ringelmann Smoke Chart. The chart, as pubHshed by 



No- '• No. 2. No. a No. 4. 

Fig. 105. Ringelmann Smoke Chart (Greatly Reduced). 

the U. S. Geological Survey and used by the Smoke Department of 
the City of Chicago and other municipahties, consists essentially of a 
cardboard folder 12 by 26 inches over all. Four charts are printed on 
this folder, each chart consisting of 294 squares, 14 squares wide by 21 
squares in length, the width of the fines and spacings varying as illus- 
trated in Fig. 105. At a distance of 50 feet from the observer the lines 
become invisible and the cards appear to be of different shades of gray, 
ranging from very fight gray to almost black. The observer places the 
chart on a level with the eye (at the distance stated, and as nearly as 
possible in line with the chimney) and notes which card most nearly 



218 



STEAM POWER PLANT ENGINEERING 



corresponds with the color of the smoke. Observations should be made 
at 15-second intervals and recorded as in Fig. 106. No smoke is re- 
corded as No. 0, 100 per cent as No. 5, and the intermediate colors as 
indicated by the cards. 

Experienced observers often record in half-chart numbers. Although 
these observations depend upon the personal element it is the opinion 
of the Chicago Smoke Department that only a little experience is neces- 
sary to effect consistent results with different observers. 




Fig. 106. Smoke Record Chart. 



Prior to 1910 a chimney was held to be a smoke nuisance by the 
Chicago smoke inspection authorities when it emitted smoke of No. 3 
density, according to the Ringelmann chart, for 7 minutes during one 
hour, as based on the original ordinance. With this standard the 
owners of a chimney which emitted but a very small total quantity of 
smoke might be Uable to punishment, whereas, with a chimney which 
continuously emitted smoke of a density less than No. 3, the owners 
would be safe from legal prosecution, although the total quantity emitted 
might be many times as great. 

The total smoke emitted is now taken into consideration. Obser- 
vations are made on a given stack every 15 seconds throughout the 
entire day and the total ''smoke units" are recorded, from which the 
average smoke density for the entire period is calculated. 

A " smoke unit " is the equivalent of No. 1 smoke (Ringelmann 
scale) emitted for one minute. No. 1 smoke has a density of 20 per 
cent; No. 2, 40; No. 3, 60; No. 4, 80; and No. 5, 100 per cent. Thus, 
if a stack emits No. 3 smoke for 6 minutes, 18 smoke units are charged 
against it. If this smoke was emitted during one hour's observation, 
then 

3 X 6 X 20 ^ 

TTTT = 6 per cent 

60 

is the average density of smoke emitted during the period of observa- 
tion. 



i 



SMOKE PREVENTION, FURNACES, STOKERS 



219 



If observations on a given stack show that the density averages 
more than 2 per cent, although the owner may not be legally hal)le, 
an appeal is made to his personal and civic pride by a representative 
of the smoke-inspection department. For example, if a certain hotel 
stack emits smoke of more than 2 per cent average density, the smoke 
department finds a plant record of similar design and equipment, 
preferably a hotel plant, which shows a record well below the 2 per 
cent mark. This plant is then pointed out to the owner or manager 
having the objectionable chimney and he is asked if he cannot do 
equally well when he has practically the same equipment, etc. 

It has been found that this method of procedure often produces quicker 
and better results than a threat to go to law. 

New Methods of Approaching the Smoke Problem: Osborne Monnett, Jour. Wes. Soc. 
Engrs., Nov. 4, 1912. 

DIVISIONS OF MESH; RINGELMANN'S SMOKE CHART. 



Numbers give 

Relative Smoke 

Density. 


Thickness of 
Lines, mm. 


Distance in the 

Clear between 

Lines, mm. 




1 

2 

3 
4 
5 


All white 

1 

2.3 

3.7 

5.5 
All black 


All white 
9.0 
7.7 
6.3 
4.5 




Fig. 107. Hammler-Fiddy Smoke Recorder — Motor-driven Type. 

The Hammler-Eddy smoke recorder, Fig. 107, is one of the most 
successful devices for automatically recording the density of the smoke 



220 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



221 







222 STEAM POWER PLANT ENGINEERING 

independent of personal observations. This apparatus consists essen- 
tially of a small motor-driven vacuum pump, which draws a continuous 
sample of the products of combustion from the uptake, breeching or 
stack and discharges it against a paper-covered drum revolved by- 
clockwork. The density of the smoke, the time at which visible smoke 
is being emitted and the duration of the smoke-production period are 
automatically recorded on the paper by the smoke itself. Before 
reaching the pumps the gases pass through a glass ''emergency" con- 
denser and a large portion of the vapor content is removed. The 
pump discharges the partially dried gases against a surface of sulphuric 
acid (which removes the last trace of moisture) and forces the smoke 
in the form of a small jet of dry powder onto the surface of the record- 
ing paper. The sampling tube leading from the flue to the pump is 
connected with a steam hne and is ''blown out" each time a card is 
changed. The instrument is very compact and portable and may be 
placed anywhere with respect to the chimney. A number of these 
appUances in Chicago power plants are giving excellent satisfaction. 
In a more recent design the pump is replaced by a steam siphon. 

106. Cost of Stokers. — The following is the relative approximate 
cost of stokers suitable for a Babcock and Wilcox boiler of 350-horse- 
power rated capacity with 45 square feet of grate surface; height of 
chimney above grate, 175 feet; coal burned, Illinois screenings. The 
cost of installation included, exclusive of brickwork, is 

1. Chain grate and appurtenances $1500 

2. Jones underfeed stoker 1400 

3. Hawley down-draft furnace 1350 

4. Burke smokeless furnace 1000 

5. Roney stoker 1300 

6. Murphy furnace and stoker 1350 

7. Wilkinson stoker 1200 

BIBLIOGRAPHY 

Boiler Settings for Smokeless Combustion: Jour. A.S.M.E., Aug. 1916, p. 633. 

Burning Soft Coal Without Smoke: Heat. & Vent. Mag., Oct., 1916, p. 19. 

Development of the Smokeless Boiler: Heat. & Vent. Mag., Oct., 1916. 

Reducing Costs with Mechanical Stokers: Eng. Mag., Nov., 1915, p. 276. 

Smoke Prevention at Boston Edison Company: Elec. Wld., Aug. 28, 1915, p. 469. 

Smoke Prevention at Chicago: Jour. Soc. Western Engrs., April, 1916, p. 310. 

Smoke Prevention at Dayton Power Plant: Power, July 25, 1916, p. 127. 

Smoke Prevention at Massachusetts: Power, Aug. 10, 1915, p. 213. 

Smoke Prevention at Pittsburgh: Power, Aug. 3, 1915, p. 152. 

Smoke Prevention at St. Louis: Elec. Wld., Jan. 22, 1916, p. 212. 

Smoke Prevention at Washington University: Assn. Eng. Soc, Nov., 1915, p. 139. 

Smoke Prevention Codes for Large and Small Cities: Heat. & Vent. Mag., Oct., 
1916, p. 25. 
Tests of Hand-fired Furnaces: Power, March 21, 1916, p. 396. 



CHAPTER V 

SUPERHEATERS 

107. Advantages of Superheating. — That superheated steam results 
in ultimate plant economy is evidenced by the fact that the largest 
and most economical plants in the world are equipped with superheaters. 
With very high pressures and temperatures, initial cost and upkeep 
may offset the thermal gain due to the use of superheated steam, but 
in general, a Hmited amount of superheat effects ultimate economy in 
nearly all cases. Practically all modern central turbo-generator stations 
and large isolated piston engine plants are designed for superheated 
steam. No general rules can be drawn as to the extent of the saving 
made because of the great number of variable factors entering into the 
problem. Each installation must be considered by itself and due con- 
sideration given to such items as the type and size of prime movers, 
character of service, nature and cost of fuel, piping, first cost, upkeep 
and the like. The logical procedure is to determine the saving in fuel 
regardless of other factors and then deduct the extra expense due to first 
cost and upkeep. The resulting net gain or loss will show whether or 
not the use of superheat is advisable. 

Theoretically, all types of steam-driven prime movers show increased 
heat efficiency with superheated steam, but the gain is usually less than 
that actually reahzed in the commercial mechanism. Aside from the 
gain in the prime mover there is the possible added efficiency in the 
boiler plant. It is true that the heat required to superheat steam is 
furnished by the fuel and when a definite weight is superheated an added 
amount of fuel must be burned, but with a properly designed super- 
heater integral with the boiler the over-all efficiency of boiler and super- 
heater is usually somewhat higher than if saturated steam alone were 
generated, so that the added amount of fuel is less than the heat gained 
by the steam. In addition to the thermal gain in the prime mover and 
boiler there may be a reduction in heat losses in the piping system be- 
cause smaller pipes may be used and because superheated steam gives 
up heat less rapidly than does wet steam. Furthermore, the increased 
economy of the prime mover may permit a reduction in the size of boilers, 
condensers and other auxiliary apparatus. 

The principal advantages of superheated steam in connection with 
piston engine work are: 

223 



224 STEAM-POWER PLANT ENGINEERING 

1. At high temperatures it behaves Uke a gas and is, therefore, in a 
far more stable condition than in the saturated form. Considerable 
heat may be abstracted without producing liquefaction, whereas the 
sUghtest absorption of heat from saturated steam results in condensa- 
tion. If superheat is high enough to supply not only the heat absorbed 
by the cylinder walls but also the heat equivalent of the work done 
during expansion, then the steam will be dry and saturated at release. 
This is the condition of maximum efficiency in a single cyhnder. (Rip- 
per, "Steam Engine Theory," p. 155.) Greater superheat than this 
may result in a loss of energy unless the steam is exhausted into another 
cylinder. To obtain dry steam at release the steam at cut-off must be 
superheated from 100 to 300 deg. fahr. above saturation temperature, 
depending upon the initial condition of the steam and the number of 
expansions, a higher degree of superheat being required for earlier cut- 
off. A superheat of from 250 to 350 deg. fahr. at admission is necessary 
to insure dry steam at release in the average single-cylinder engine 
cutting off at one-fourth stroke, boiler pressure 100 pounds gauge. In 
most cases superheat is only carried so far as to reduce initial conden- 
sation, the steam becoming saturated at cut-off, thus permitting efficient 
lubrication. There will be a reduction of approximately 1 per cent in 
cylinder condensation for every 7.5 to 10 degrees of superheat. In 
compound and triple-expansion engines the steam is ordinarily super- 
heated between each stage as well as before admission to the high- 
pressure cylinder. 

2. A moderate amount of superheat produces a large increase in 
volume, the pressure remaining constant, and diminishes the weight 
of steam per stroke for a given amount of work. For example, the 
volume of one pound of saturated steam at 165 pounds pressure (abso- 
lute) is 2.75 cubic feet, and its temperature is 366 deg. fahr. The 
total heat of one pound of this steam above the freezing point is 1195 
B.t.u. By adding 108 B.t.u. in the form of superheat its temperature 
will be increased to 565.8 deg. fahr. (superheated 200 deg. fahr.) and its 
volume to 3.68 cubic feet (specific heat taken as 0.54). Thus an in- 
crease of 9 per cent in the heat effects an increase of 34 per cent in the 
volume, which means a corresponding reduction in the weight of steam 
admitted to the engine per stroke. These figures are purely theoretical, 
as no allowances have been made for condensation of the saturated 
steam or for reduction in temperature of the superheated steam. 

3. Superheated steam has a much lower thermal conductivity than 
saturated steam, and, therefore, less heat is absorbed per unit of time 
by the cyhnder walls. 

The water rate of the steam turbine is decreased by superheating 



SUPERHEATERS 225 

but to a less extent than the piston engine. Theoretically the improve- 
ment in steam economy is the same for both types of prime movers, 
pressure and temperature ranges being the same in each case, but in 
actual practice the gain is more pronounced with the piston engine. 
In general, the less economical the steam motor the more is the gain 
effected by superheating. Aside from the gain in heat efficiency the 
use of superheated steam benefits the turbine by reducing erosion of 
the blades and by lowering skin friction and windage. The fact that 
nearly all modern steam turbine plants are operated with superheated 
steam is evidence that superheating results in ultimate plant economy. 

108. Economy of Superheat. — Many comparative tests of engines 
and turbines using saturated and superheated steam under varying 
conditions of pressure and temperature have been made during the 
past few years, showing in all cases decreased steam consumption due 
to superheat. In the majority of moderately superheated steam in- 
stallations the ultimate gain was a substantial one, but in a few cases 
involving the use of very -high temperatures and pressures, the extra 
investment and cost of maintenance neutralized the reduction in steam 
consumption, resulting in an actual loss when measured in dollars and 
cents per unit output. With high degree of superheat (over 250 deg. fahr.) 
apparatus of a special nature is necessary and it is questionable whether 
the additional first cost, care and habiUty to operating difficulties, up- 
keep and maintenance will not offset any fuel saving accompUshed. 

As far as steam consumption per horsepower-hour is concerned, super- 
heating usually increases the economy of the piston engine from 5 to 15 
per cent and in some instances as much as 40, the latter figure referring 
to the more wasteful types. A fair estimate of the average reduction 
in steam consumption per horsepower-hour with moderate superheating, 
that is, from 100 to 125 deg. fahr., based on continuous operation of 
existing plants, is: p^^Cent. 

1. Slow running, full stroke, or throttling engines, including direct- 

acting pumps 40 

2. Simple engines, non-condensing, with medium piston speed, in- 

cluding compound direct-acting pumps 20 

3. Compound condensing Corliss engines 10 

4. Triple-expansion engines 6 

European builders guarantee steam consumption with highly super- 
heated steam (total temperatures 750 to 850 deg. fahr.) as follows: 

Pounds per 
I.hp-hour 

Single-cylinder condensing engines (uniflow) 8.5 

Single-cylinder non-condensing engines (uniflow) 12.0 

Compound condensing engines (locomobile) 8.0 

Compound non-condensing engines (locomobile) 10 . 5 



226 STEAM POWER PLANT ENGINEERING 

An exceptionally low steam consumption is credited to a locomobile 
compound using steam superheated to 806 deg. fahr. at an initial 
pressure of 220 pounds absolute. When exhausting against an ab- 
solute back pressure of 1.32 pounds the steam consumption was 6.95 
pounds per i.hp-hour. (Zeit. des Ver. deut. Ingr., Mar. 18, 1911, p. 415.) 

In high-pressure steam turbines the water rate is improved approxi- 
mately one per cent for every 8 to 12 deg. fahr. superheat; the higher 
rate holding for about 50 degrees superheat and the lower for about 
200 degrees. It is difficult to estimate the actual gain in heat economy 
due to superheating in very large turbines, since they are not designed 
for saturated steam and tests with the latter do not offer a true compar- 
ison. In a general way the average reduction in steam consumption 
for these large units is about 1 per cent for every 10 deg. fahr. in- 
crease in superheat. One of the best recorded performances is that of 
a 20,000-kilowatt turbo-generator installed in the New River station 
of the Buffalo General Electric Co.; with initial absolute pressure of 
265 lb. per sq. in., 275 deg. fahr. superheat and absolute back pressure 
of 1 inch of mercury, the steam consumption was 10.25 lb. per kilowatt- 
hour. 

In comparing the performances of engine and turbines using saturated 
steam it is advisable to base all results on the heat consumed per unit 
output rather than on the steam consumption, since the latter is apt 
to give a false idea of the relative economies. The real measure of 
economy is the cost of producing power, taking into consideration all 
charges, fixed and operating, and the next best is the coal consumption 
per unit output, but as a means of comparing the motors -only, the 
heat consumption per unit output is very satisfactory. (See para- 
graph 162.) 

See paragraph 182 for the influence of superheat on the economy of 
reciprocating engines and paragraph 221 for the influence on steam 
turbines. 

109. Limit of Superlieat. — In this country steam temperatures ex- 
ceeding 600 deg. fahr. are seldom employed, while in Europe few if 
any plants are installed without superheaters, and 600 degrees is a 
common temperature with a maximum of about 850. There is no 
particular mechanical difficulty in designing power plant apparatus 
to withstand temperatures as high as 850 deg. fahr., and for industrial 
purposes steam temperatures of 1000 deg. fahr. are not uncommon, but 
first cost and maintenance usually offset any thermal gain accomplished 
except perhaps where fuel is very high and labor is cheap. In this 
country where fuel is comparatively cheap but material and labor are 
high, a moderate amount of superheat appears to effect the best economy. 



SUPERHEATERS 227 

Experience has shown that with engines of ordinary design, shde- 
valves and Corliss, the temperature at the throttle should not exceed 
500 deg. fahr. This corresponds to a superheat of 160 degrees with 
steam at 100 pounds gauge pressure, and 130 degrees at 150 pounds. 
This degree of superheat insures practically dry steam at cut-ofT in the 
better grade of engines. Just how far superheating can be carried with 
a given engine of ordinary construction can be determined by experiment 
only, but a temperature of 500 degrees is probably an outside figure 
and 450 degrees a good average. Higher temperatures are apt to inter- 
fere with lubrication and sometimes cause warping of the valves. With 
temperatures below 450 degrees no difficulties are ordinarily met with. 

With highly superheated steam involving temperatures of 600 deg. 
fahr. or more, the poppet-valve type of engine (Figs. 196, 203) is or- 
dinarily employed, though balanced piston and specially designed 
Corliss valves are not uncommon. The poppet valve is not distorted 
by heat and requires no lubrication. In Europe these engines have 
been brought to a high state of efficiency, but have not been generally 
adopted in this country. The steam end of the composite gas-steam 
engines at the Ford Motor Company's plant, Detroit, are of Corliss 
valve design and though the steam at admission has a temperature of 
700 deg. fahr., no difficulty is experienced with lubrication. 

Owing to the absence of rubbing parts in contact with the steam, 
and because the casing is not subjected alternately to high and low 
temperatures, steam turbines may be designed to operate successfully 
with temperatures up to 800 deg. fahr., though temperatures above 
600 deg. are exceptional. The majority of turbine installations in 
this country, including the very latest, are designed for temperatures 
under 650 degrees. 

Properties of Superheated Steam. — See Chapter XXII. 

How to Use Superheated Steam: Eng. Mag., May, 1916, p. 208; June, 1016, p. 413. 

110. Types of Superheaters. — Superheaters are manufactured by 
practically all boiler builders, the characteristics of the boiler being 
embodied to a large extent in the design of the superheater. The 
superheater may be independently fired or placed in the boiler setting. 
In the latter arrangement the superheater may be located in the furnace, 
as in Fig. 53, at the end of the heating surface as in Fig. 116, or at some 
intermediate point, as in Figs. 55 and 110. Since the absorption of 
heat depends chiefly upon the average temperature difference between 
the gases and the steam and the extent of superheating surface, the 
required degree of superheat may be obtained from a small extent of 
heating surface in the furnace, a large amount in the rear of the heating 



228 • STEAM POWER PLANT ENGINEERING 

surface or a proportionate amount in intermediate locations. In a 
general sense the sum of the boiler heating surface and superheating 
surface per boiler horsepower is practically the same for any degree of 
superheat. The cost of a superheated steam boiler is approximately 
equal to that of a saturated steam boiler since the superheated plant has 
less steam to generate. The requirements of a successful superheater 
are: 

1. Security of operation, or minimum danger of overheating. 

2. Economical use of heat appKed. 

3. Provision for free expansion. 

4. Disposition so that it may be cut out without interfering with 
the operation of the plant. 

5. Provision for keeping tubes free from soot and scale. 

Superheaters may be separately fired or indirectly fired. The advan- 
tages of the separately fired superheater are: 

1. The degree of superheat may be varied independently of the per- 
formance of the boiler. 

2. It may be placed at any desired point. 

3. Repairs are readily made without shutting down the boiler. 

Some of the disadvantages are: 

1. It requires separate attention. 

2. Saturated steam only can be furnished to the prime movers in 
case of a breakdown to the superheater. 

3. Extra piping is required. 

4. Extra space is required. 

The indirectly fired superheater arranged in the boiler setting has 
the advantage of: 

1. Lower first cost. 

2. Higher operating efficiency. 

3. Minimum attention. 

4. Minimum space requirements. 

As ordinarily installed the indirectly fired superheater^ is 'subject 
to the fluctuating temperatures of the furnace so that forcing the boiler 
has a similar effect on the superheater. In some cases the superheater 
adjusts itself automatically to the load requirements maintaining a 
constant degree of superheat at all loads, but in most cases the degree 
of superheat increases with the load, see Fig. 124. Standard central 
station practice in this country favors the superheater contained within 
the boiler setting. 

Figs. 110 and 111 show the application of superheating coils to a 



SUPERHEATERS 



229 




Fiu. 110. Babcofk and Wilcox Superheater. 




Fig. 111. Babcock and Wilcox Superheater. 



230 STEAM POWER PLANT ENGINEERING 

Babcock and Wilcox boiler illustrating the usual location of the in- 
directly fired type. The superheater consists of two transverse square 
wrought steel manifolds into which two sets of 2-inch cold drawn seam- 
less steel tubes bent to a ^'U" shape are expanded. The tubes ordi- 
narily are arranged in groups of four. Saturated steam flows from the 
dry pipe located within the drums to the upper manifold. The latter 
is divided into as many sections as there are drums so as to avoid 
expansion strain. From the upper manifold the steam passes through 
the ''U" shaped tubes to the lower one (which is continuous) and thence 
to a cast-steel ''superheater center" fitting supported over the drum. 
The "superheater center" fitting is provided with a superheated steam 
outlet and an extra opening for the reception of the superheater safety 
valve. This safety valve is furnished as a part of the regular equip- 
ment and is set two pounds lower than the safety valves of the boiler. 
This is essential so as to provide a flow of steam through the superheater 
and to prevent any overheating of the latter in case the load should be 
suddenly thrown off the boiler. A small pipe connects the center 
fitting with the saturated steam space in the drum and is for the pur- 
pose of equalizing the pressure when the discharge from the superheater 
is closed. While a flooding device is not necessary its use is recom- 
mended by the Babcock & Wilcox Company. This consists essentially 
of a small pipe connecting the lower manifold with the water space of 
the boiler and by means of which the superheater may be flooded. 
Any steam formed in the superheater tubes is returned to the boiler 
drum through the collecting pipe which, when the superheater is at 
work, conveys saturated steam into the upper manifold. When steam 
pressure has been attained the superheater is thrown into action by 
draining the water away from the manifolds and opening the superheater 
stop valve. With the proportion of superheating surface to boiler sur- 
face ordinarily adapted the steam is superheated from 100 to 150 deg. fahr. 




Fig. 112. Section Through Superheater Header and Tubes showing Method of Hold- 
ing Core in Place: Babcock & Wilcox Superheater as Applied to Stirling Boiler. 

When the boiler construction permits of only one inlet and one outlet 
connection to the superheater the Babcock and Wilcox superheater 
is modified by using one set of ''U" tubes fitted with cores. Such 
a modified type is used in connection with the StirUng boiler. The 



SUPERHEATERS 



231 



cores are made of No. 13 B.W.G. tubes plugged at one end and inserted 
in the straight portion of the 2-inch superheater tubes, thereby causing 
the steam to flow through the annular space. Fig. 112 shows a cross 
section through an element of the superheater header and tubes, illus- 
trating the method of holding the core in place. 

Fig. 113 shows the appUcation of a Foster superheater to a Babcock 
and Wilcox boiler. This device consists of cast-iron headers joined by 
a bank of straight parallel seamless drawn-steel tubes, each tube being 
encased in a series of annular flanges placed close to each other and 




Fig. 113. Foster Superheater in Babcock and Wilcox Boiler. 



forming an external cast-iron covering of large surface. The protection 
afforded by this external covering is ample to prevent damage from over- 
heating during the process of steam raising, and flooding devices are 
unnecessary. The tubes are double, the inner tube serving to form a 
thin annular space through which the steam passes as indicated. Caps 
are provided at the end of each element for inspection and cleaning 
purposes. Foster superheaters are more costly than plain-tube super- 
heaters, but are longer Hved and offer a much larger heating surface 
in proportion to the space occupied. 
The ''Schwoerer" superheater, which is somewhat similar in external 



232 



STEAM POWER PLANT ENGINEERING 





§■ 



SUPERHEATERS 233 

appearance to the Foster, differs from it considerably in detail, the 
heating surface being made up of suitable lengths of cast-iron pipe 
ribbed outside circumferentially and inside longitudinally. The ends 
of the pipes are flanged and connected by cast-iron U-bends. The 
intention is to pro^^de ample heating surface internally and externally, 
with a compact apparatus. 

Fig. 114 shows the application of a Heine superheater to a Heine 
boiler, illustrating the installation of a superheater within the boiler 
setting but entirely separated from the main gas passages. The super- 
heater consists essentially of a number of IJ-inch seamless steel tubes, 
bent to U-shape and expanded into a header box of the same type 
of construction as the standard Heine boiler water leg. The interior 
of this box is divided into three compartments by light sheet-iron dia- 
phragms, so as to deflect the current of steam through the tubes. The 
superheater chamber is located above the steam drum as indicated. 
The gases of combustion are led to the superheater chamber through a 
small flue built in the side walls of the setting. A damper placed at 
the outlet of the flue controls the flow of gases and regulates the degree 
of superheat. No provision is necessary for flooding the superheating 
coils since the gases may be entirely diverted from the heating surface. 
Soot accumulations are readily removed by introducing a soot blower 
through the hollow stay bolts. 

The Schmidt independently-fired superheater, Fig. 115, consists of 
two nests of coils, A and D, of equal size and dimensions, connected 
to cast-iron headers and I. Saturated steam enters the first nest 
of coils through C and passes into header 0. From the steam, which 
is now dried, and partly superheated, flows through the cast-iron pipe 
E to header /, and thence through the second nest of coils into header 
adjoining 0, and through pipe R to the engine. In chamber D the 
steam and gases flow on the counter-current and in chamber A on the 
concurrent principle. This combination permits of a low flue tempera- 
ture and high steam temperature without subjecting the tubes to an 
excess of heat as would be the case if the steam left the coils A at header 
7, where the furnace gases are the hottest. A steam temperature of 
750 deg. fahr. and a flue temperature of 450 deg. fahr. are easily main- 
tained with this apparatus. A mercury pyrometer T is fitted where 
the superheated steam enters the discharge pipe R. A thermometer 
cup L permits of checking the pyrometer by means of a nitrogen-filled 
thermometer. Each coil can be taken out separately and a new one 
put in without removing the others or dismantling the plant. Water 
produced by condensation while the superheater is inoperative collects 
in the bottom header N and escapes through a drain cock. If the steam 



234 



STEAM POWER PLANT ENGINEERING 




SUPERHEATERS 



235 




L^'-.v. Av.E£ Y^''-'^''-' '^ 



Fig. 116. Schmidt System of Combined Superheater, Feed-water 
Heater and Economizer. 



Inlet 




Fig. 117. Foster Independently-fired Superheater. 



236 STEAM POWER PLANT ENGINEERING 

supply should be suddenly shut off, the air door P is opened automati- 
cally by weight K. As soon as steam begins to flow it raises the weight 
through the opening of valve C and the door closes. The Schmidt 
superheater when arranged in the flue has practically the same construc- 
tion as the independently fired. 

Modern Superheater and its Performance: Ry. Age Gaz., June 30, 1916. 

111. Materials Used in Construction of Superheaters. — Most super- 
heaters are constructed either of wrought steel, cast iron, or cast steel, 
the latter material having the advantage of not being damaged by any 
temperature to which it is Hkely to be subjected, which does away with 
the necessity of damper mechanisms and simplifies the installation. 
Cast-metal superheaters are usually ribbed after the fashion of an air- 
cooled gas engine, and are, therefore, very heavy and thick walled, 
necessitating a higher temperature for the same useful effect than in 
the case of the wrought-iron construction, but have the advantage of 
minimizing fluctuation of steam temperature which would otherwise be 
caused by a wide variation in temperature of f.urnace. One of the most 
successful cast-metal heaters is of European design and is constructed of 
a special alloy known as ^'Schwoerer" iron. Table 42 gives the yearly 
cost of repairs to piping and necessary brickwork for a number of in- 
stallations equipped with cast-metal superheaters of the '^Schwoerer" 
type. 

Wrought steel offers the advantage of lightness, ease of construction, 
and low first cost, but cannot be exposed to very high temperatures 
without injury, and consequently provision must be made for diverting 
the direction of the heated gases or for flooding the coils while the 
boiler is being warmed before steam is generated. 

The effect of temperature on superheater materials is shown in Fig. 
118. It will be seen that the tensile strength drops off very rapidly for 
temperatures beyond 600 deg. fahr. Because of this rapid decrease 
in tensile strength of materials with the increase in temperature, steam 
is seldom superheated to temperatures above 850 deg. fahr. 

For further information pertaining to the effect of temperature on 
various metals, consult ''The Effect of High Temperatures on the Physi- 
cal Properties of Some Metals and Alloys"; The Valve World, Jan., 
1913, published by the Crane Co., Chicago. 

Ordinary cast-iron valves and fittings have shown permanent in- 
crease in dimensions under high superheat and in numerous instances 
have failed altogether, but sufficient data are not available to prove 
conclusively the unreUability of cast iron if the iron mixture is properly 
compounded and the necessary provision is made for expansion and 
contraction. Authorities are of the opinion that the failure of cast-iron 



SUPERHEATERS 



237 



u 

>8 
< 


For 1 Foot 
of Ele- 
ment. 


d 

CO 


CO 

o 

CO 


00 

o o 

o 


o 
v.- 


00 

o 
o 


o 


oo 
d 


^1 


o 


o 
o 


o 

Oi 

^ o 


CO 

CO 

00 
CO 




o 




Length of 
One Ele- 
ment in 
Feet. 


o 


CO 

oo 

05 


<N CO 
O Tt< 
Cq OO 

00 d 


CO 

QO 
05 


«o 








No. of 

Elements 
for One 
Super- 
heater. 


1—1 


■* 


(M to 


«3 


o 
o 

00 






d 


Average 
Temp, of 

Steam. 
Degrees F. 


So 








o 




o' 
c^ 

o> 

CO 




oo 


o 
o^ 




CO 

Oi 


Average Steam 

Pressure. 

Pounds per Square 

Inch 

(Absolute). 




o 


CO 05 
00 CO 


CO 


«o 




CO 


Average Temp, 
of Gases im- 
mediately in 
Front of the 
Superheater 

Surface. 
Degrees F. 










■ o 

. "^ 

. oo 
•o' 

; CO 


















Place of Installation of 
Superheater with Ref- 
erence to Boiler. 


1 

> 

1 




1 

C 

h 




f a. Directly behind the . 

firebridge 

' 6. Behind the first flue . 


i 

t 

a 

c 




i 
Is 

ll 




1 

c 
Q 


1 














"""^ 




Average 

Daily 

Use. 

Hours. 


- 


(M 


- 


C<1 




- 


CO 


Length of 
Time of 
Installa- 
tion. 
Years. 


oo 


ITS 


o 


- 


to 


b- 






1 
S 

3 






(N 




CO 




rt 




U5 




<£> 




1 


2 



238 



STEAM POWER PLANT ENGINEERING 



fittings is due more to fluctuations in temperature than to the actual 
high temperature itself and cite numerous cases where ordinary cast- 
iron fittings under uniform temperature conditions are giving satis- 
faction with highly superheated steam. Notwithstanding the claims 



400 450 500 



rOO 750 800 



1000 




300 400 450 500 600 700 750 800 

Temperature, Degrees Fahrenheit 

Fig. 118. Effect of Temperature on Strength of Materials. 



1000 



that cast iron properly compounded is a perfectly reliable metal for 
fittings, engineers are inchned to use cast or forged steel, at least in this 
country. See ''Effect of Superheated Steam on Cast Iron and Steel," 
Trans. A.S.M.E., Vol. 31, 1909, p. 989. 

112. Extent of Superheating Surface. — The required extent of super- 
heating surface for any proposed installation depends upon: (1) the 
degree of superheat to be maintained; (2) the velocity of the steam 
and gases through the superheater; (3) the character of the superheater; 
(4) the weight of steam to be superheated; (5) the moisture in the wet 



SUPERHEATERS 239 

steam; (6) the temperature of the gases entering and leaving the super- 
heater; (7) the conductivity of the material, and (8) cleanUness of the 
tubes. 

Since the heat absorbed by the steam in the superheater is equal to 
that given up by the products of combustion, neglecting radiation, this 
relationship may be expressed 

SUd = Wc{h-t2), (53) 

in which 

S = square feet of superheating surface per boiler horsepower, 

U = mean coefficient of heat transmission, B.t.u. per hour per 

degree difference in temperature, 
d = mean temperature difference between the steam and heated 

gases, deg. fahr., 
W = weight of gases passing through the superheater per boiler 
horsepower-hour, 
c = mean specific heat of the gases, 

ti = temperature of the gases entering superheater, deg. fahr., 
tz = temperature of the gases leaving superheater, deg. fahr. 

Transposing equation (53), 

The heat transfer from the products of combustion to the steam may 
also be expressed 

SUd = wc' {ts - t), (55) 

in which 

w = weight of steam passing through the superheater, pounds per 

boiler horsepower-hour, 
c' = mean specific heat of the superheated steam, 
ts = temperature of the superheated steam, deg. fahr., 
t = temperature of the saturated steam, deg. fahr., 
S, U, and d as in equation (53). 
For wrought-iron or mild steel tubes U varies as follows: 
U = 1 to 3 for superheaters located at the end of the heating surface 
= 3 to 5 for superheaters located between the first and second pass 

kof water tube boilers 
= 8 to 12 for superheaters located immediately above the fur- 
nace in stationary boilers, in the smoke box of locomotive 
boilers, and in separately fired furnaces. 

General practice allows i to ^ square foot of superheating surface per 
boiler horsepower for mild steel superheaters located in the furnace; 
from 2 to 2.5 square feet of surface at the end of the first pass, and from 



240 STEAM POWER PLANT ENGINEERING 

3 to 4 square feet at the end of the heating surface for superheats of 
from 100 to 150 deg. fahr., boiler pressure 150 pounds absolute. 

The Foster Superheater Company allows 6 B.t.u. per Uneal foot per 
degree difference in temperature for their ' 'two-inch " element where 
the average temperature of the gases is about twice the mean temper- 
ature of the steam. 

For all engineering purposes d may be determined with sufficient ac- 
curacy from the relationship 

, _ ^i_±l2 _ is + t ^ 
"^ ~ 2 2 * 

Notations as in equations (53) and (55). 

An empirical formula for determining the extent of superheating 
surface in connection with indirect superheaters which appears to con- 
form with practice is derived by substituting 

U = 3, d = t' - ^-^\ w = 30, c' = 0.5, 

in equation (55) [J. E. Bell, Trans. A.S.M.E., 29-267]. Thus: 

>S X 3 (^' - ^^) = 30 X 0.5 X {ts - t), 



from which 



^'S^^r « 



t' (the mean temperature of the product of combustion where the super- 
heater is located) may be approximated from equation 

y_\^o.i6 = 0-172 H + 0.294, (57) 

in which 
H = the per cent of boiler-heating surface between the point at 

which the temperature is t and the furnace, 
t as in (56). 

Equation (57) is based upon the assumption that the heat trans- 
ferred from the gases to the water is directly proportional to the differ- 
ence in temperature; that the furnace temperature is 2500 deg. fahr.; 
flue temperature 500 deg. fahr.; steam pressure 175 pounds per square 
inch gauge; one boiler horsepower is equivalent to 10 square feet of 
water-heating surface; and that there is no heat absorbed by direct 
radiation. 

Example 15. What extent of heating surface is necessary to superheat 
saturated steam at 175 pounds gauge pressure, 200 deg. fahr., if the 

* See also paragraph 286. 



SUPERHEATERS 



241 




SaSV9 A8 H3A9 OaSSVd aSVJUns 9NUV3H UXLVM JO XN3a Hid 



242 



STEAM POWER PLANT ENGINEERING 



superheater is placed in the boiler setting where the gases have already 
traversed 40 per cent of the water-heating surface? 
Substitute H = 0.4 and t = 378 in equation (57), 



from which 



(f - 378)016 



0.172 X 0.4 + 0.294, 



f = 950. 



Substitute i' 



950, ts -- 

S = 



578, and t = 378 in equation (56), 
10 (578 - 378) 



2 X 950 - 578 - 
= 2.12 square feet. 



378 



The curve in Fig. 119 was plotted from equation (57) and gives a 
ready means of determining t' and of observing the law governing heat 
absorption by the boiler between furnace and breeching. The abscissas 
represent the temperatures of the hot gases at different points in their 
path between furnace and breeching. The ordinates represent (1) the 
per cent of boiler-heating surface passed over by the hot gases, and 
(2) the per cent of the total heat generated which is absorbed by this 
heating surface. 

In the use of equation (57) the probability of error is greatest when 
considering a point near the furnace, since large quantities of heat are 

transmitted to the tubes by radiation 
from the fuel bed which are not taken 
account of. For most practical pur- 
poses the assumption is sufficiently 
accurate. 

Fig. 120 gives the probable temper- 
ature range of gases entering super- 
heater after passing over a given per 
cent of boiler-heating surface and 
Fig. 121 shows the relation between 
lo superheating surfaces and boiler heat- 
ing surface. (See Power, Nov. 7, 

1911, p. 696.) 

Fig. 120. Temperature Range of Gases j^ ^-^y ^^ f^^^^ ^^^^ ^^^ boiler- 

in Superheater. , ,- r i -i i 

heatmg surface per boiler horsepower 

will be decreased in almost the same proportion that the superheating 

surface is increased, so that the sum of the boiler-heating surface and 

superheating surface per boiler horsepower will be very nearly the same 

for any degree of superheat. 

113. Performance of Superheaters. — Published tests of both directly 

and indirectly fired superheaters cover such a wide range of conditions 




60 55 50 

Per Cent 

Boiler Heating Surface Used 

befo£e_Ileaching Superheater 



SUPERHEATERS 



243 



Fig. 

















J 650 

Ic^coo 

^0.550 
||500 












y 










/ 


/ 








/ 


/' 






y 


/ 








/ 


/ 










/ 













10 



15 



35 40 



25 30 
Per Cent 
(Superheater Surface in Per Cent 
of Boiler Heating Surface) 

121. Relation between Superheat 
and Boiler Heating Surface. 



^% 




















l'S20 
II .0 


































^ 
















^ 
























Percen 
Prodi 

























200 400 600 800 

Horse Power Produced in Boiler 
Fig. 122. Ratio of Horsepower Produced 
in the Superheater to that Developed in 
the Boiler. 



900 






































1 








800 
















■ 






















/ 








700 














































600 














































S 500 












'"■'■■'i 






















4 












n 












f 
























/• 












300 










/ 


1 ■ 




















/ 














200 










/ 




















•/ 
















100 








/. 




















/ 




















--r: 


y 





















50 100 150 200 250 

Degrees in Superheat— F 
Fig. 123. Relation of Degree of Super- 
heat to Total Horsepower Developed. 





80 


Ui 




a 




1. 


70 




60 










^ 


50 






^ 




S 


40 


w 




n 


30 


tt 





10 



7 



200 



" 50 100 150 

Degrees of Superheat— F 
Fig. 124. Relation of Degree of Sui)erheat 
to Horsepower of Superheater. 



244 STEAM POWER PLANT ENGINEERING 

of installation and operation that general conclusions cannot be drawn, 
but it may be of interest to note briefly the performances in a few 
specific cases. 

The curves in Figs. 122, 123, and 124 are plotted from tests of a Bab- 
cock and Wilcox boiler, with 5000 square feet of water-heating surface, 
equipped with superheating coils of 1000 square feet area, as illustrated 
in Fig. 93. The furnace with ordinary short ignition arch was pro- 
vided with chain grate of 75 square feet area. 

Fig. 122 shows the relation between degrees of superheating and total 
horsepower of boiler and superheater. 

Fig. 123 shows the relation between horsepower produced in the 
boiler and the percentage of boiler horsepower produced in the super- 
heater. 

Fig. 124 shows the relation between the degree of superheat obtained 
and the horsepower developed in the superheater. 

Tables 42 to 45 are taken from the report of Otto Berner (''Zeit. d. 
^'er. Deut. Eng. " and reprinted in Power, August, 1904). 

Table 42 compares the heat efficiency of a steam plant equipped with 
indirectly and with separately fired superheaters, the former showing a 
much higher efficiency. 

Table 43 compares different boilers with and without flue super- 
heaters, showing the effect upon the temperature of the flue gases. 
The gain in heat efficiency of the entire plant due to the use of the 
superheater is decisive in each case. 

Table 45 shows the gain in heat efl&ciency due to the use of super- 
heaters in a number of plants equipped with fire-tube boilers. 

Table 46 gives the results of tests on one of the return tubular boilers 
at the Spring Creek Pumping Station of the Brooklyn Waterworks 
(Feb. 9, 1904) with and without a superheater. The superheater, of 
the Foster type, was installed between the rear wall of the setting and 
the tube sheet. 

Although the results in Tables 42 to 46 represent practice of ten years 
ago, they agree substantially with current practice (1916). 



SUPERHEATERS 



245 



ai 



(- 


tL 


s. 


1 


M 


x: 


aj 


05 


3 


,^ 


fa 


C 


a 


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02 



Oi 



W to 

^ 02 



c3 

o 



CO OS c^ 
05 (M lO 


00 lO O 
OO ■* CO o ■«*< 




Oi C^ Oi Oi 
(N (M CO 
00 <M 


<N ■* «0 O O 




rH CO 







-* oo 

OS C^l 


"■f CO ^ O 

O Oi 05 O CO • 


CO 

CO 


OO 


CO CO -^ 

S2 




(M 

05 
CO 


CO 


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<M 





CO C^ Oi 05 
CO (M CO 

CO (M 



C<|-*»OC00t^-<*ilO 
r- 1 t^ CO O oo -"^ O 

T-H lO T-C »0 CO 



COCOCOrt* »— i05t^-^t<COCOcOO 

CO TfO t^coooiot^oo 

CO OOrH T-1 T*<100000 



c3 
02 



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248 



STEAM POWER PLANT ENGINEERING 



TABLE 46. 
{Engineer, U.S., May 1, 1904.) 



Time of start. 
Time of finish 



Hours run 

Average steam pressure 

Average water pressure, triple expansion, 

head in feet 

Average water pressure, compound, head 

in feet 

Average vacuum of suction for triple and 

compound, inches of mercury 

Total head on triple, feet of water 

Total head on compound, feet of water . . . 

Total double strokes, triple 

Total double strokes, compound 

Gallons pumped from piston displacement, 

total, triple 

Gallons pumped from piston displacement, 

total, compound 

Gallons pumped from piston displacemen!:, 

total, triple combined 

Gallons, total, pumped as measured by weir 

Per cent slip 

Foot pounds, weir 

Total coal consumed 

Per cent refuse 

Total refuse 

Total feed water 

Duty per 100 pounds coal 

Duty per 1,000 pounds steam 



With Superheater. 



12 noon, Feb. 8 
12 noon, Feb. 9 



24 
79.3 1b. 

0.99 

7.10 

22.90 
29.05 
33.04 
30,557 
35,395 

2,854,023 

2,930,706 



5,784,720 


5,848,930 


4,492,680 


4,549,480 


22.3 


22.2 


,163,815,819 


1,184,983,596 


5,015 1b. 


6,410 lb. 


23.7 


18.7 


1,188 


1,203 


38,399 


50,960 


23,206,696 


18,486,483 


30,308,498 


23,253,213 



Without Superheater 



11 A.M., Feb. 11 
11 A.M., Feb. 12 



24 
79.4 1b. 

1.05 

7.10 

23.21 
29.46 
33.39 
34,114 
32,158 

3,186,247 

2,662,682 



Per cent increase of work per 100 pounds coal 25.5 

Per cent increase of work per 1000 pounds steam 30.2 

Per cent saving in coal per foot pound work 20.2 

Per cent saving in feed water per foot pound work 23.2 

Average temperature steam leaving superheater 527.4 deg. fahr. 

Average temperature steam entering superheater 320. 1 deg. fahr. 

Average degree superheat 207 . 3 deg. fahr. 



PROBLEMS. 

1. Required the sq. ft. of superheating surface necessary to superheat 10,000 lb. 
of saturated steam per hour at 200 lb. abs., to 250 deg. fahr. if the superheater is 
placed in the boiler setting where the gases have already traversed 35 per cent of the 
water-heating surface. 

2. Required the mean temperature of the products of combustion passing through 
a superheater of 3815 sq. ft. of heating surface if 66,000 lb. of steam are superheated 
from saturation at 265 lb. abs. to 250 deg. fahr.; lb. gas per lb. of steam, 2.0; mean 
coefficient of heat transfer, 5. 

3. If the furnace gases enter a superheater at a temperature of 1200 deg. fahr. and 
leave at 900 deg. fahr., required the weight of steam superheated from saturation 
at 265 lb. abs. to 250 deg. fahr.; lb. gas per lb. of steam 2.0. Neglect all losses. 



CHAPTER VI 

COAL AND ASH-HANDLING APPARATUS 

114. GeneraL — The cost of coal and its delivery into the furnace 
are usually the largest items in the operating charges; hence largo 
central stations are located, when practicable, adjacent to a railway 
hne or water front, to minimize the cost of handling coal and ashes. 
Isolated stations in the business districts of large cities are usually 
unfavorably situated, so that the cost of handling coal and ashes is a 
large percentage of the total fuel cost. In large stations the amount 
of fuel and ash handled frequently warrants the expense of elaborate 
conveyor systems which would not be justified in smaller plants. In 
whatever way coal is supplied pro\ision should be made for storing a 
quantity sufficient to operate the plant for some time in case the supply 
is interrupted, thereby guarding against an enforced shut-down. 

If adjacent to a railway line, a side track must be provided for switch- 
ing the cars. As the furnishing of bottom-dumping cars cannot be 
depended upon, provision should be made for unloading by hand or 
by grab bucket. If coal is dehvered by water, clam-shell drop buckets 
are ordinarily used for unloading the barges. If the power house is 
located at some distance from the railroad or water the coal is generally 
hauled by teams or motor trucks in two- to five-ton loads. 

115. Coal Storage. — In small stations the storage bins or coal 
bunkers may usually be located within the building, but in larger plants 
the quantity of coal consumed daily is frequently such that an immense 
space would be required to furnish storage capacity for even a short 
period of time. For example, the requirements of the Northwest Sta- 
tion of the Commonwealth Edison Company are approximately 3000 tons 
of Illinois coal per day of 24 hours. One day's supply would occupy a 
space 100 feet square and 12 feet high. Seventy-five railway cars per 
day would be required to supply this amount of coal and in addition 
about ten cars of ashes would have to be removed. The futility of stor- 
ing the coal in cars is evidenced by the fact that about two and one-half 
miles of track would be necessary to carry only a four days' supply. In 
this particular plant there is yard space for storing 300,000 tons in two 
piles, or sufficient to run the plant for three months. 

Exposed coal piles are objectionable, because of freezing in winter, 
the crust sometimes freezing so hard as to necessitate the use of dyna- 

249 



250 STEAM POWER PLANT ENGINEERING 

mite to break it; moreover, a slow depreciation in heat value takes 
place, especially with bituminous coal. This depreciation is more 
rapid in warm weather and in the tropics. Stored coal is oftentimes 
subject to spontaneous combustion, particularly when there is a large 
content of iron pyrites. Storage under water minimizes spontaneous 
combustion and depreciation in heat value. (Consult references below.) 

Coal bunkers or hoppers are ordinarily placed on the same level with 
the boiler-room floor or above the boiler setting. The former is the 
cheaper as far as the first cost is concerned, but necessitates additional 
handling of the fuel before it can be fed to the stokers. In the overhead 
system the coal gravitates to the stoker through down spouts. Over- 
head bunkers are usually found where real estate is costly. They are 
generally constructed of steel plates lined with concrete or of reenforced 
concrete. The bottoms slope at an angle of 35 to 45 degrees and empty 
into the coal chutes or down spouts. Fig. 130 shows the general ap- 
pearance of a steel plate overhead bunker and Fig. 136 that of the sus- 
pended type. In some bunkers the floors are made with very slight 
slopes, but it is not advisable to use a slope less than the angle of repose 
of the coal, as it may be necessary to shovel the coal over the spouts. 
Convenience in framing makes the 45-degree slope the more desirable. 
Separate bunkers for each boiler are preferred to continuous bunkers, 
since fire in the coal is more readily prevented from spreading. In 
the new power house of Swift & Co., Chicago, 111., the bunkers are of 
circular cross section instead of rectangular, as is the usual practice. 
The capacity of the cylindrical hopper is considerably less than that 
of a rectangular hopper of the same width, but is much cheaper to' 
construct. 

Ash bins are invariably lined with concrete or brickwork, since the 
corrosive action of the ashes would soon destroy the bare iron, and are 
usually located as in Fig. 130 so that they may be discharged by gravity. 
The angle of repose of most ashes is approximately 40 degrees, but the 
45-degree angle is preferred on account of convenience in construction. 

Coal Storage Under Water: Elec. Wld., Oct. 7, 1911, p. 885; Eng. News, Dec. 24, 
1908; El. Ry. Jour., June 24, 1916, p. 1191. 

Calorific Value of Weathered Coals: Bulletin No. 17, Univ. of 111., Aug. 26, 1907; 
Eng. News, Jan. 11, 1912, p. 64. 

Spontaneous Combustion of Coal: Jl. Ind. and Chem. Eng., Mar., 1911. 

Suspended Coal Bins: Power, Apr. 23, 1912, p. 602. 

Concrete Coal Cylinders: Eng. News, Mar. 2, 1916, p. 420. 

116. Coal Handling Methods. — The best method of dehvering coal 
to the furnace and of removing refuse from the ash pit is the one which 
will effect the desired result at the lowest ultimate cost. That this 



i 



COAL AND ASH-HANDLING APPARATUS 251 

problem does not offer a simple solution is evidenced by the almost 
endless combinations found in practice for the same operating conditions. 
The principal factors which influence the choice of system are size and 
location of plant and cost of fuel and labor. In pubhc service plants 
continuity of operation may be of even greater importance than re- 
duction of cost and extra investment may be considered advisable to 
offset the unreliable labor element. Of the various methods found in 
current practice the following are the more common: 

1. Hand shoveling. 

2. Wheelbarrow or hand car and shovel. 

3. Continuous conveyors: 

Spiral or screw, ""^^^ 

Scraper or flights, . 
Apron and buckets, 
Overlapping pivoted buckets, 
Endless belt. 

4. Hoist and hand car. 

5. Hoist and automatic cable car. 

6. Hoist and trolley: telpherage. 

7. Clam shell buckets. 

8. Vacuum system. 

9. Combinations of the above. 

117. Hand Shoveling. — Where possible, the coal is dumped direct 
from the cars or wagons into bins located in front of the boilers. In 
such instances one man may handle the coal and ashes and attend to 
the water level of 200 horsepower of boilers equipped with common 
hand-fired furnaces. With hand-shaking and dumping grates 300 
horsepower may be fired by one man and from 800 to 1000 horsepower 
with automatic stokers. This refers, of course, to average good coal 
not too high in ash nor productive of much chnker. Sometimes the 
coal cannot be stored in front of the boilers but must be hauled by 
wheelbarrow, cart, or rail car. For distances over 100 feet and quanti- 
ties over 20 tons per day the cost of handhng the coal in this way may 
justify the installation of an automatic conveyor system. Hand-fired 
furnaces and manual handling of coal and ashes are usually associated 
with small plants of 500 horsepower and under, but a number of large 
stations are operated in this way with apparent economy. A notable 
example is the steam power pknt of the Wood Worsted Mill, Lawrence, 
Mass., in which 40 return tubular boilers are fired by hand. A tipcart 
with a capacity of one ton brings the coal a distance of 100 to 200 feet 
to the firing floor, and firemen shovel it on to the grate. Four men 
are stationed at the coal pile. One man drives two carts (one of which 



252 



STEAM POWER PLANT ENGINEERING 



is being filled while the other is gone with its load), sixteen firemen attend 
to the furnaces, and two men dispose of the ashes. 

Most large plants, however, are equipped with conveying machinery, 
not so much because of the possible reduction in cost of operation, tak- 
ing into consideration all charges fixed and operating, as because of the 
large and often unreliable labor staff which it dispenses with. Hand 
shoveling is sometimes necessary even with modern unloading devices 
on account of the freezing of coal in the cars. This is particularly true 
of washed coals, and it is not unusual to have an entire car load solidly 
frozen. In this case it has to be picked and shoveled by hand, or the 
unloading tracks must be equipped with steam pipes and outfits for 
thawing purposes. A good man is capable of shoveling 40 to 50 tons of 
coal in eight hours when unloading a car, provided it is only necessary 
to shovel the 'coal overboard. For cost of handling material by wheel- 
barrow and hand shoveling see end of paragraph 123. 

118. Continuous Conveyors and Elevators. — The most popular method 
of automatically handhng coal and ashes in the modern power plant 
is by means of continuous conveyors and elevators. They may be 
divided into two general classes: 

1. Those which push or pull the material, but in which the weight 
of the load is not borne by the moving parts. 

2. Those which actually carry the load. 

A few of the more important types will be treated briefly. 

Screw or Spiral Conveyors. — These consist of a stamped or rolled 
sheet steel spiral secured by lugs to a hollow shaft (usually a standard 
or extra heavy pipe) revolving in a trough which it fits approximately. 
Standard sizes range from 3 to 18 inches in diameter and in sections 
from 8 to 12 feet long. 



TABLE 47. 

SPEEDS AND CAPACITIES OF SCREW CONVEYORS. 
(Fine Coal and Ashes.) 



Diam. screw, in 

Max. r.p.m 

Capacity per hr., fine coal, tons. 
Ashes, cu. ft 



6 
115 


7 
110 

7 
175 


8 

105 

14 

350 


9 

100 

16 

425 


10 

95 

21 

550 


12 

90 

36 

950 


14 

85 

48 

1200 


16 

80 

80 

2000 


125 



18 
75 

no 

2700 



On account of the torsional strain on the shaft the maximum length 
seldom exceeds 100 feet. Single sections may be used as feeders on 
incUnes up to 15 degrees. Low first cost, compactness and adaptability 
to space requirements are the advantages of this type but these may be 
offset by high power consumption and excessive wear. The following 



COAL AND ASH-HANDLING APPARATUS 



253 



equation gives a means of approximating the power requirements for 

horizontal runs : 

WL 
Horsepower = C (58) 

in which 

C = 0.7 for coal and LO for ashes, 
W = capacity in lb. per minute, 
L = length in feet. 

Fig. 125 shows an application of a screw conveyor for handling coal 
as installed in a modern isolated station. 




t'lCC-fi; I Bri^nchTunL^el^ Illinois Tuuucl^C^uip^aDy, ^ < 




Screw Conveyor as Installed in the Power Plant of a Tall Office 
Building. 

Scraper or Flight Conveyor. This consists of a trough of any desired 
cross section and a single or double strand of chain carrying flights 
or scrapers of approximately the same shape as the trough. The 
flights scrape the material along the trough discharging at the end of 
trough gate controlled openings in the bottom of the conduit. Three 
types of flight conveyors are in common use; plain scraper, suspended 
flights and roller flight. In the plain scraper the flights are suspended 
from the chain and drag along the bottom of the trough. In the sus- 
pended flight conveyor the flights are attached to cross bars having wear- 
ing-shoes at each end and do not touch the trough at any point. The 



254 



STEAM POWER PLANT ENGINEERING 



roller flight differs from the suspended type only in the substitution of 
rollers for the wearing-shoes. A typical installation of scraper and 
roller flight conveyors is illustrated in Fig. 126. The coal conveyor is 
a single strand roller flight, 80 feet in length between centers, driven by 
a 5-hp. electric motor. It has a capacity of 15 tons of buckwheat coal 
per hour. The ash conveyor is a single-strand drag-chain with 87 ft. 
centers on the horizontal run and 6 ft. between vertical centers. The 




Fig. 126. Scraper and Drag Conveyor as Installed in a Power House of the Otis 

Elevator Company. 

chain operates in an extra heavy cast-iron trough set in a cement trench 
and is operated by a 5-hp. motor. Fhght conveyors are low priced and 
offer an economical and efficient means of handling coal and ashes in 
small plants. 

Ajyron Conveyors are commonly used for conveying coal from track 
hopper to the main conveyor and elevator. The most elementary 
form consists of flat steel plates attached between two chains and form- 
ing a continuous platform or apron. Since the load is carried and not 



1 



COAL AND ASH-HANDLING APPARATUS 



255 



smwwwwxw^ 



!#H#^- 







256 



STEAM POWER PLANT ENGINEERING 



dragged less power is required than with the scraper type and the main- 
tenance is lower. These carriers are not suitable for elevating material 
except at an inclination not exceeding 30 degrees. End discharge only 
is possible. Fig. 127 shows a typical apron conveyor installation. 

Pan Conveyors and Open Top Conveyors are similar to the apron carriers 
except that pans or buckets take the place of the flat or corrugated 
apron plates. These conveyors are used where pans deeper than those 
of an apron conveyor are required, as on inclines too flat for elevators 




Fig. 128. Typical Installation of V-Bucket Conveyor for Handling Coal and 
Cast-iron Pan Conveyor for Handling Ashes. 

and too steep for efficient operation of flight or apron conveyors. Con- 
veyors of this type are usually run at speeds of 30 ft. to 50 ft. per min- 
ute and when equipped with self -oiling roflers of 6-inch to 8-inch diameter 
demand but little power for their operation above theoretical load 
requirements. Fig. 128 shows an installation of a cast-iron pan conveyor 
for handling ashes. 

The power required to operate flight, apron and open top conveyors 
may be closely approximated by the following empirical equation.* 



Hp. = 



AWLS . BLT 



1000 



+ 



1000 



+ x 



(59) 



C. K. Baldwin, The Robins Conveying Belt Co. 



COAL AND ASH-HANDLING APPARATUS 



257 



in which 

Hp. = the horsepower required at the conveyor drive shaft, 
A, B = constants as in Table 48, 

W = weight of conveyor per ft. of run, lb., 
L = distance between centers of head and tail sprockets, ft., 
S = speed of conveyor, ft. per min., 
T = capacity of conveyor, tons (2000 lb.) per hour, 
a; = 1 for conveyors up to 100 ft. centers and 2 for longer con- 
veyors. 
If the convej^or is composed of portions on different inclines compute 
the power for each section separately and add 10 per cent for each change 
in direction. 

The V-Bucket Conveyor consists of a series of V-shaped buckets 
rigidly fastened to the conveyor chain. The buckets act essentially 
as a drag conveyor on horizontal runs, each bucket pushing its half- 
spilled load ahead of it through a suitable trough. On vertical runs 
they act as elevators. A typical V-bucket conveyor for handling coal 
and a pan conveyor for handhng ashes are illustrated in Fig. 128. The 
power requirements may be approximated from the following empirical 



equations ; 



^ AWL'S BUT TH 



1000 



1000 ' 1000 



in which 

L' = horizontal length of conveyor, ft., 

Li = total horizontal length traversed by the loaded bucket, ft., 

H = total vertical traverse, ft., 

x' = number of 90-degree turns in the conveyor. 

Other notations as in equation (59). * 



(60) 



TABLE 48. 

VALUES OF CONSTANTS IN CHAIN CONVEYOR POWER FORMULAS. 



Angle of 
Conveyor 
with Hori- 
zontal Deg. 


A. 


B. 

Scraper, Apron and 

Open Top. 


B. 
V-Bucket and Pivoted 
. Bucket. 


Sliding 
Block. 


3^in. 

Roller, 

3Hn. 

Pin. 


6-in. 

Roller, 

l|-in. 

Pin. 


6-in. 
Roller, 


Anthra- 
cite 
Coal. 


Bitu- 
minous 
Coal. 


Ashes. 


3^in. 

Roller, 

?-in. 

Pin. 


6-in. 

Roller, 

l|-in. 

Pin. 


6-in. 
Roller, 

'ft 





0.030 


0.0043 


0.0046 


0.0050 


0.33 


0.60 


0.54 


0.071 


0.076 


0.083 


6 


0.030 


0.0043 


0.0046 


0.0050 


0.43 


0.69 


0.63 


0.17 


0.18 


0.19 


12 


0.030 


0.0042 


0.0045 


0.0049 


0.54 


0.79 


0.73 


0.28 


0.28 


0.29 


18 


0.029 


0.0041 


0.0044 


0.0048 


0.63 


0.88 


0.82 


0.38 


0.38 


0.39 


24 


0.028 


0.0039 


0.0042 


0.0046 


0.72 


0.95 


0.90 


0.48 


0.48 


0.49 


30 


0.026 


0.0037 


0.0040 


0.0043 


0.79 


1.02 


0.97 


0.57 


0.57 


0.58 


36 


0.025 


0.0035 


0.0037 


0.0040 


0.86 


l.OS 


1.03 


0.65 


0.66 


0.66 


42 


0.023 


0.0032 


0.0034 


0.0037 


0.92 


1.12 


1.07 


0.73 


0.73 


0.74 


48 


0.020 


0.0029 


0.0031 


0.0033 


0.97 


1.15 


1.11 


0.80 


0.80 


0.81 



258 



STEAM POWER PLANT ENGINEERING 




COAL AND ASH-HANDLING APPARATUS 



259 




Fig. 130. Coal and Ash-handling System in the Power House of the South Side 
Elevated Railway Company, Chicago. 



260 



STEAM POWER PLANT ENGINEERING 



The Pivoted Overlapping Bucket Conveyor is perhaps the most popular 
type of continuous conveyor for large power plant service. It consists 
essentially of a continuous series of buckets pivotally suspended between 
two endless chains. The buckets at all times maintain their carrying 
position by gravity whether the chain is horizontal, vertical or inclined. 
By means of this system no transfer of material is necessary and dis- 
charge may be made at any desired point. Fig. 129 gives a diagram- 
matic arrangement of the Peck Carrier illustrating the principles of 
a complete coal and ash-handling system and Fig. 130 illustrates its 
apphcation to a typical boiler plant. 

Coal is discharged from the railway cars into a track hopper and from 
there delivered by a '''feeding apron" into a crusher which reduces it 



rp;^^^ 




Fig. 131. Crusher and Cross Conveyor at the Power Plant of the South Side 
^iu Elevated Station, Chicago. 



to such a size as can be conveniently handled by the stokers. It is 
then discharged into a short apron or pan conveyor, which carries it 
to the main system of buckets, and it is elevated to the proper level and' 
discharged into the overhead bunkers. The discharge is effected by 
special tripping devices which engage the buckets and turn them over. 
The ashes are dumped from the ash pit through a series of chutes into 
the lower run of buckets, by which they are elevated and discharged 
into the ash hopper alongside the coal bunkers. From the ash hopper 
the ashes discharge by gravity directly into the railway cars below. 
The system is operated by means of two motors, one driving the crusher 
and the other the main bucket system. The buckets are made of 
malleable iron. 

In Fig. 129 the coal is fed to the crusher by the "reciprocating feeder," 
which is usually placed directly under the track hopper. The feeder 



COAL AND ASH-HANDLING APPARATUS 



261 



consists of a heavy steel plate mounted on rollers and having a recipro- 
cating movement effected by a crank mechanism from the carrier. 
The amount of coal deUvered depends upon the distance the plate 
moves, and this can be varied by changing the throw of the eccentric. 






Coal Conveyors- ^ 




Fig. 132. Coal and Ash-handling System of the Commonwealth Edison 
Company, ''Northwest Station." 

The number of strokes corresponds to the number of buckets. Any 
size coal can be readily handled. When the distance from track hop- 
per to carrier is so great that the reciprocating feeder is not practi- 
cable, a continuous or "belt" feeder is used to supply the crusher with 



262 



STEAM POWER PLANT ENGINEERING 



fuel. The ''equalizing gear" is designed to impart a pulsating motion 
to the driving sprocket wheel which will counteract the natural pulsa- 
tion to which long pitch chains are subject, producing violent increase 
of the normal strain at frequent intervals. This is accompHshed by 
driving the spur wheel with an eccentric pinion, causing the pitch Hne 
to describe a series of undulations corresponding to the number of 
sprockets on the chain wheel. Figs. 130 and 131 show the general 
arrangement of crusher and ''cross conveyor" in the old portion of the 
South Side Elevated Power House, Chicago. 

A coal and ash system similar to the one illustrated in Fig 129 for a 
plant consisting of eight 350-horsepower boilers will cost in the neigh- 
borhood of $8000, completely installed. This does not include the 
cost of coal and ash bunkers. 




Fig. 133. Driving Mechanism of Hunt Conveyor. 

The Hunt Conveijor, Fig. 133, while usually called a "bucket" con- 
veyor, is in fact a series of cars connected by a chain, each having a 
body hung on pivots and kept in an upright position by gravity. The 
chain is driven by pawls instead of by sprocket wheels. The "buckets" 
are upright in all positions of the chain, consequently the chain can be 
driven in any direction. The change of direction of the chain is ac- 
complished by guiding the carriers over curved tracks. The chain 
moves slowly, and the capacity is governed by the size of the buckets. 
The ordinary size buckets carry two cubic feet of coal and move at a 
rate of fifteen buckets a minute, carrying about 40 tons per hour. 
Two methods of filhng the buckets are employed, the "measuring" 



COAL AND ASH-HANDLING APPARATUS 



263 



and the ''spout filler." In the former each bucket is separately filled 
with a predetermined amount by a suitable ''measuring feeder." In 
the latter the material is spouted in a continuous stream, necessitating 
the use of overlapping buckets to prevent spilling of the material. 
Fig. 134 shows an application of the Hunt system to the power plant 
of the Rhode Island Suburban Railway, Providence, R. I. 

The power required to operate carrier conveyors of the pivoted 
bucket type may be approximated from formula (60), using the proper 
value for B as given in Table 48. 




Fig. 134. Coal and Ash-handling System, Rhode Island Power House. 



Belt Conveyors have a distinctive advantage over most other types 
of carriers in that they may be driven from any point in their length. 
The driving machinery is extremely simple; power is appUed to one 
or more pulleys over which the conveyor belt passes. The maximum 
width of conveyors is Kmited only by the fiber stress in the belt. Con- 
veyors 1000 feet from center to center handling 500 tons per hour 
have been successfully operated. Inchnations are limited by the angle 
of repose of the material. In power plant service they seldom exceed 
20 degrees. 



264 



STEAM POWER PLANT ENGINEERING 



ua 




Fig. 135. 



Guide Pulleys, Robins Belt 
Conveyor. 



The Robins Belt Conveyor, Fig. 135, consists essentially of a thick 
belt of the required width driven by suitable pulleys and carried upon 
idlers so arranged that the belt becomes trough-shaped in cross section. 
For heavy duty five pulleys are employed instead of three as illustrated 
in order that the line of contact may more nearly approach the arc 
of a circle. The belt is constructed of woven cotton duck covered 
with a special rubber compound on the carrying side. The rubber is 
thicker at the middle than at the edges, since the wear is greatest in 
a line along the center, but the thickness of the belt is uniform through- 
out its entire width. The edges are reenforced with extra piles of duck 

to increase the tensile strength. 
The idlers are carried by iron 
or wooden framework, and are 
spaced from 3 to 6 feet between 
centers on the troughing side, 
according to the width of belt 
and the weight of the load. On 
the return side these distances 
range from 8 to 12 feet. High- 
speed rotary brushes with inter- 
changeable steel bristles prevent wet, sticky material from chnging to 
the belt. Automatic tripping devices placed at the proper points cause 
the material to be discharged where it is needed. The trippers consist 
essentially of two pulleys, one above and shghtly in advance of the 
other, the belt running over the upper and under the lower one, the 
course of the belt resembling the letter S. The material is discharged 
into chutes on the first downward turn of the belt. The trippers may be 
movable or fixed, single or in series. Movable trippers are used when it 
is desired to discharge the load evenly along the entire length, as, for 
instance, in a continuous row of bins, while fixed trippers are employed 
where the load is to be discharged at certain and somewhat separated 
points. The movable trippers are made in two forms, '^hand-driven" 
and '' automatic." In the former they are moved from point to point 
by means of a hand crank. The ''automatic" tripper is propelled 
by the conveying belt through the medium of gearing. It reverses 
its direction automatically at either end of the run and travels back 
and forth continuously distributing its load. It can be stopped, re- 
versed, or made stationary at will. Notable installations of this 
system are at the Hudson and Manhattan Railway Company's power 
house, Jersey City; L Street Station, Edison Illuminating Company 
of Boston; South Boston Power Station of the Boston Elevated Com- 
pany and the Essex Power Station of the Public Service Electric Co., N. J. 



COAL AND ASH-HANDLING APPARATUS 



265 




.s 

a 






266 



STEAM POWER PLANT ENGINEERING 



The power required to drive belt conveyors may be approximated from 
the following empirical equation. (C. K. Baldwin, Trans. A.S.M.E., 
Vol. 30, p. 187.) 

For level conveyors: 

Hp. = ^- (61) 



For incHned conveyors: 



Hp. 



1000 



CTL , TH 



+ 



1000 ' 1000 



(62) 



C = constant as given in Table 49, 

T = load in tons (2000 lb.) per hr., 

L = length of conveyor between centers, ft., 

H = vertical lift of material. 

For each movable or fixed tripper add the horsepower given in Table 49. 
For friction of conveyor ends and driver add the following: 

Length of conveyor 25 50 75 100 200 500 

Per cent added power 80 50 30 20 10 4 



TABLE 49. 

POWER REQUIREMENTS FOR BELT CONVEYORS. 
(Coal and Ashes.) 



Width of belt 

C 

Hp. required for each movable 
or fixed tripper 



12 
0.234 

0.50 



16 
0.220 

0.75 



20 
0.205 

1.25 



24 
0.195 

1.5 



28 
0.175 

2.25 



32 
0.163 

2.75 



36 
0.157 

3.25 



Belt-conveyor Operating Data: Power, Oct. 3, 1916, p. 490. Economics of Conveyor Equipments: Eng. 
Mag., Nov., 1916, p. 231. 

119. Elevating Tower, Hand-car Distribution. — Fig. 137 illustrates 
the coal and ash-handhng system as originally installed at the Aurora 
and Elgin Interurban Railroad power house, Batavia, 111. Coal is 
delivered to the plant by railroad cars which dump directly into coal 
hoppers located inside a steel structure running the entire length of 
the building and spanned by two railroad tracks. There are 18 hoppers 
constructed of 17-inch brick walls fitted with steel-plate bottoms. 
Subdividing the storage space in this manner makes it possible to carry 
different grades of coal, prevents the spreading of fire, and affords a 
simple construction for the support of the railroad tracks. The base- 
ment of the boiler room extends underneath the hoppers, and two 
lines of narrow-gauge tracks are embedded in the concrete floor. Turn- 
tables at the center facilitate the switching of cars to the elevators 
which rise through the boiler room close to the chimney. The cars, 



COAL AND ASH-HANDLING APPARATUS 



267 



of one ton capacity each, are of special construction, with roller-bearing 
axles and a combined ratchet lift and friction dump. The filled cars 
are pushed from underneath the hoppers to two elevators which lift 
them to the hne of tracks supported overhead across the boiler fronts. 



iiiffiiSs^aiii;^!! 




Track to Elevator 



Fig. 137. Typical Coal and Ash-handling System Involving the Use of Elevating 
Tower and Hand-car Distribution. 



They are then pushed to the hoppers suspended above the boiler set- 
ting and the coal is dumped. These hoppers have a capacity of six 
tons each. From the hoppers the coal is fed to the stoker by an 
ordinary down spout. The ashes fall from the stokers into an ash 
pit, from which they may be discharged into ash cars. The ash cars 
are elevated to a set of tracks running at right angles to the main 



268 STEAM POWER PLANT ENGINEERING 

tracks, and are transferred to ash bins located directly over the coal 
bins. Coal and ashes are weighed in the small cars. There are ten 
boilers in this plant and four men are required to handle the coal and 
ashes. The entire coal and ash-handUng system cost about $10,000, 
and the cost of handUng the coal and ashes, exclusive of fixed charges, 
is approximately 4 cents per ton. This does not include wages of 
firemen or water tenders. For a description of recent changes made 
in this plant see Prac. Engr., U. S., Nov. 1, 1916, p. 907. 

120. Elevating Tower, Cable-car Distribution. — The coal and ash- 
handhng system of the Delray Station of the Detroit Edison Company 
is a typical example of a large station equipped with elevating tower 
and cable-car distributers instead of the usual bucket conveyor. The 
system consists essentially of a lofty steel tower in which are housed 
at various levels a track receiving hopper, crushing rolls and feeders, 
weighing hopper, hoisting apparatus, etc., and a small cable railway 
for delivery to the bunkers. The railroad coal cars enter the tower 
on an elevated trestle 18 feet above grade, below which is a track 
receiving hopper. A two-ton "tub hoist" is filled with coal from the 
bottom of the receiving hopper and elevated to a 20-ton bin at the top, 
120 feet above ground level. This bin has a grille bottom at one side 
and under the outlet a heavy duty coal crusher, thus allowing the 
fine coal to screen through directly while all the larger lumps are auto- 
matically deHvered to the crusher. From the two bins the small cable 
cars are filled for dumping into the desired bunkers over the boiler 
rooms. The cars are arranged for automatic dumping by means of 
adjustable trips which may be located at any point. The entire sys- 
tem has a capacity of from 125 to 150 tons of coal per hour and is 
motor-driven. The ash-handling system consists of brick-lined con- 
crete hoppers underneath each pair of stokers which discharge their 
contents by gravity into the small cars operated on the track system 
in the boiler-house basement. 

When handUng 600 tons per day of 24 hours the cost of operation is 
approximately 20 cents per ton from coal car to ash car. This includes 
wages of firemen and water tenders. 

131. Hoist and Trolley; Telpherage. — The telpher is a form of 
electric hoist which lifts and transfers the load on overhead tracks 
from one point to another. Fig. 138 illustrates a very simple and 
economical method of handhng coal and ashes as installed by the 
Jeffrey Manufacturing Company at the power plant of the Scioto 
Traction Company embodying the telpher systems. If the coal car is 
of the dump type the contents are discharged directly into the coal 
pit from which the coal is removed by grab bucket and transferred 



COAL AND ASH-HANDLING APPARATUS 



269 







s 



270 STEAM POWER PLANT ENGINEERING 

either to the overhead bunker or to the storage pile. If the coal car 
is of the gondola type the coal is removed directly from the car by 
the grab bucket. The bucket is hoisted and carried on the trolley 
into the building over the screen hoppers where it discharges its con- 
tents; the finer particles fall directly into the bunker and the larger 
lumps are automatically dehvered to the crusher. The grab bucket 
will take about 98 per cent of the coal in the car, leaving only 2 per 
cent to be handled by hand. Coal is fed to the stokers by means of 
a traveUng electric hopper which receives its supply from the over- 
head bunkers. The present capacity of the plant is 50 tons per hour 
taken from the car or pit to stock pile. 

132. Vacuum Conveyors. — This type of conveyor is finding favor 
with many engineers for handhng dry ashes because of its simphcity 
in design and ease of application. It has also been used in a few cases 
for handling small nut coal and screenings. The system consists 
essentially of a pipe line through which air is flowing at a high ve- 
locity. The material to be conveyed is fed into the pipe through 
suitable openings and the momentum of the column of air carries it 
to the point of discharge. Velocity is imparted to the air either by 
a mechanical exhauster or by steam jets discharging in the direction 
of flow. 

Fig. 139 gives a diagrammatic arrangement of a vacuum ash-con- 
veying system as installed in the power plant of the Armour Glue 
Works, Chicago, Illinois. One end of special cast-iron header F leads 
to the ash pits of the various boilers by means of branch tubes, and 
the other end is connected with the closed storage tank. Each branch 
pipe is fitted with simple circular openings directly underneath each 
ash-pit door for admitting ashes. These openings are kept covered 
except when in operation. Exhauster E creates a partial vacuum in 
chamber D and draws in air at a high velocity from the opening in the 
ends of the branch pipes. Ashes raked into the pipes through the 
openings are caught by the rapidly moving column of air and forced 
into the storage tank. Air is withdrawn from the top of the sepa- 
rator chamber through pipe G and discharged to the stack or to waste. 
A spray is introduced into pipe F to reduce dust. In this particular 
installation the system is applied to a boiler plant of thirteen boilers, 
aggregating 4800 horsepower, and cost, completely installed, $5600. 
The ash bin has a capacity of 60,000 pounds of wet ashes and is con- 
structed of five-sixteenths-inch sheet iron. The exhauster (a 30-foot 
Root blower) has a capacity of about 8000 cubic feet per minute at 
265 r.p.m., and is driven by a 75-horsepower motor. Under normal 
conditions of operation the motor requires 50 horsepower when de- 



COAL AND ASH-HANDLING APPARATUS 



271 




272 



STEAM POWER PLANT ENGINEERING 



livering 250 pounds of ash per minute, and the vacuum on the suction 
side of the exhauster is 3.3 inches of mercury. The pipe from the 
ash bins to the separating chamber is 10 inches in diameter and is 
constructed of number 16 and number 20 galvanized iron. The ashes 
are raked by hand from the ash pits to the suction openings of the 
branch pipes, and are handled dry, the dust being taken along with 
the ashes. Short elbows are soon worn out by the abrasive action 
of the ashes, and tees are used instead, since the accumulation in the 
"dead" end receives the impact and takes up the wear. Long radius 
bends may be used in place of the tees. The cost of power for 
handling the ashes in this installation is approximately 7 cents per ton. 




Fig. 140. Vacuum System for Handling Coal and Ashes at the Plant of the 
Pierce-Arrow Motor Car Co. 



This type of vacuum system is used for handling coal at the power 
house of the Pierce-Arrow Motor Car Company, Buffalo, New York, 
and is giving satisfactory results. The steam jet system, however, 
is used for handhng the ashes. It differs from the mechanical ex- 
hauster type in that steam jets are used for creating the vacuum. 
The jets are inserted in the pipe line between the last boiler and the 
point of delivery and discharge in the direction of flow. In the Pierce- 
Arrow plant the labor cost of handhng the coal and ashes is 20.6 cents 
per ton on a basis of 26,000 tons per year. The entire equipment 
cost $34,000. 

In the steam jet ash system, two conveying pipe sizes are in common 
use; one with 6-inch inside diameter for capacities up to three tons 
per hour and one with 8-inch inside diameter pipe for capacities from 
three to eight tons per hour. Larger sizes have not proved practi- 



COAL AND ASH-HANDLING APPARATUS 273 

cable because of the excessive amount of steam required to effect the 
desired result. For horizontal runs under 100 feet in length one jet 
placed in the elbow of the riser is sufficient to move the material, but 
for longer runs additional jets in the horizontal pipe are necessary. 
The jets as ordinarily installed require from 175 to 275 lb. of steam 
per ton of ashes per hour depending upon initial pressure and quaUty 
of the steam and size of pipe. 

Vacuum Ash-Removal System: Power, April 7, 1914, p. 473; Jan. 13, 1914, p. 41. 

123. Cost of Handling Coal and Ashes. — In large stations where 
a number of men are employed to handle coal and ashes only it is a 
simple matter to divide the cost of handling into the various stages, thus: 

1. Cost of unloading cars or barges. 

2. Cost of conveying coal to bunkers. 

3. Cost of feeding coal to furnace. 

4. Cost of removing ashes. 

These costs are usually expressed in cents or dollars per ton of coal 
burned, or in terms of cents or dollars per horsepower-hour or kilo- 
watt-hour of main prime-mover output. Item number 3 is oftentimes 
included under ''boiler-room attendance" and items 1, 2, and 4 under 
''coal and ash handling." Not infrequently all four items are included 
under "attendance." So much depends upon the character of stokers 
and furnace, size of boilers, and the like, that general figures on the cost 
of handling the coal and ashes are of little value unless accompanied by 
a description of the equipment. For the sake of general comparison 
the most satisfactory method of expressing the cost is in dollars per ton 
of coal from coal car to ash car. This includes wages of coal and ash 
passers, repair men, and boiler tenders. In small stations the coal 
and ash handling is done by the boiler tenders, in which case it is im- 
practicable to separate the items mentioned above, and the cost is 
ordinarily included under attendance. An average figure for handling 
coal by barrow and shovel is not far from 1.6 cents per ton per yard 
up to the distance of five yards, then about 0.1 cent per ton per yard 
for each additional yard. With automatic conveyors the operating 
cost, not including wages of firemen and water tenders, varies with the 
size of plant and the type of conveyor, and ranges anywhere from a frac- 
tion of a cent per ton to four or five cents per ton. The larger the plant 
and the greater the amount of coal handled the lower will be the cost 
per ton. In comparing the relative costs of manual and automatic 
handling, fixed charges of at least 15 per cent of the first cost of the 
mechanical equipment should be charged against the latter in addition 
to the cost of operation. In large central stations equipped with stokers 



I 



274 



STEAM POWER PLANT ENGINEERING 



and conveyors and consuming 200 tons or more of coal in twenty-four 
hours, the cost of handhng the coal from coal car to ash car, including 
wages of firemen and water tenders but exclusive of fixed charges, will 
range between 18 cents and 25 cents a ton. 

134. Coal Hoppers. — Fig. 141 shows a front and side elevation of 
a typical set of stationary weighing hoppers as applied to the boilers 
of the Quincy Point power plant of the Old Colony Street Railway 




Fig. 141. Stationary Coal Weighing Hoppers. 



Company, Quincy Point, Mass. Each battery of boilers is provided 
with an independent set of hoppers. The bottoms of the overhead 
coal bunkers lead into the small hoppers A, A. The operation of any 
single weighing hopper is as follows: Coal is fed from the overhead 
bunkers to weighing hopper H by means of valve V. The weight of 
coal in the weighing hopper is transmitted by a system of levers and 
knife edges to the inclosed scale beam / and noted in the usual way. 
The weighed charge of coal is then admitted to the down spout S by 
means of valves similar to those at F. 



COAL AND ASH-HANDLING APPARATUS 



275 



Although separate weighing hoppers for each battery, as illustrated 
in Fig. 141, offer many advantages, they are quite costly and it is not 
unusual to install one or more large weighing hoppers mounted on 
overhead traveling carriages so that one may supply a number of boilers 
(Fig. 142). At the Armour Glue Works, Chicago, the coal supply is 
stored in one large overhead bunker of 1000 tons' capacity. A five-ton 
motor-driven traveUng hopper receives its supply from this central 




Fig. 142. Traveling Coal Hoppers. 



bunker and delivers it to the various boilers. One man operates the 
traveUng hopper, tends to the coal valves, and supplies all boilers 
with coal. 

Weighing hoppers are sometimes made automatic; that is, the open- 
ing and closing of valves, feeding of coal, and recording of weight are 
automatically performed by the weight of the coal itself. The scale is 
set for discharges of a certain weight and continues to discharge this 
amount automatically. In the few plants which are equipped with 
automatic weighing hoppers the capacity of the hopper is approximately 
100 pounds per discharge. These hoppers are necessarily more com- 
plicated and more costly than the ordinary weighing hoppers, and it is 
a question whether the advantages offset the extra first cost and main- 
tenance charges. A small automatic hopper of 100 pounds discharge 
capacity costs approximately $400 as against S250 for the ordinary 
weighing device. For a description of a coal meter see paragraph 396. 



276 



STEAM POWER PLANT ENGINEERING 



125. Coal Valves. — Figs. 145 to 147 illustrate the principles of a 
few well-known coal valves. They may be conveniently grouped into 
two classes according to the location of the coal pocket: (1) those 




Fig. 143. Common Slide Coal Valve. 



Fig. 144. Simplex Coal Valve. 



drawing the coal from overhead bunkers and (2) those drawing from the 
side of a bin. In the first class come the simple slide valve and the sim- 
plex and duplex rotating valve. In the latter are the flap valve and the 
rotating valve. They are made in various sizes and designs, but those 
illustrated are examples of the most common types. The simple slide 

valve, Fig. 143, is appHcable only 
to small size coal and to small 
spouts, since coarse or lump coal 
may get in the way and pre- 
vent proper closing. The simplex 
valve. Fig. 144, consists of a ro- 
tating jaw actuated by a lever. 
There are no rubbing surfaces, 
and the jaws cut through the 
material without jamming. The 
duplex valve. Fig. 145, consists of 
two rotating jaws connected to a 
common actuating lever. The 
jaws move simultaneously, so that 
even a partially open valve de- 
hvers the coal centrally. When 
closing the valve the flow is gradually stopped by the decreasing width 
of the opening and there is but Httle resistance to the movement of the 
jaws. The largest valve can easily be operated by hand. 




Fig. 145. Duplex Coal Valve. 



i 



COAL AND ASH-HANDLING APPARATUS 



277 



The flap valve, Fig. 146, is the simplest form for drawing coal from a 
side bin. It consists merely of an iron flap hinged to the bottom of the 
chute. The valve is lowered to let the coal run over its top and is raised 
to stop the flow. It cannot be clogged or get jammed in closing. The 
flap is raised and lowered by a simple lever. For very large bins, where 
the valves are to be opened and closed frequently, the "Seaton" valve. 
Fig. 147, is usually preferred. This valve consists of two jaws EE', 
and TT' pivoted to suitable framework at and actuated by lever A. 




Fig. 



146. Common 
Coal Valve. 



Flap" 



Fig. 147. "Seaton" Coal Valve. 



The valve is shown fully closed. Raising lever A causes the cut-off 
blade EE' to rotate about and permits the coal to flow through the 
space between the edge of the jaw E and the end of the chute. The 
rate of flow is regulated by the width of this opening. The cut-off 
blade does not reach a stop, hence there is no possil)ihty of a lump of 
coal getting in the way and preventing the prompt closing of the valve. 



BIBLIOGRAPHY. 

Coal and Ash Handling System at Connors Creek, Detroit Edison: Jour. A.S.M.E., 
Sept., 1915, p. 499. 

Coal and Ash Handling System at Delray, Detroit Edison: Aug;. 31, 19L5, p. 286. 

Coal and Ash Handling System at Essex Plant, N. J. Public Service: Prac. Engr., 
Sept. 15, 1916, p. 771. 



278 STEAM POWER PLANT ENGINEERING 

Coal and Ash Handling System at Grundy Plant, Bristol, Pa.: Power, Oct. 3, 1916, 
p. 480. 

Coal and Ash Handling System at Northern Ohio Traction Co.: Power, Sept. 21, 1915, 
p. 398. 

Coal and Ash Handling System at Northwest Station, Commonwealth Edison Company: 
Power, May 30, 1916, p. 769. 

Coal and Ash Handling System at Pacific Mills, South Lawrence, Mass.: Prac. Engr., 
Oct. 1, 1913, p. 973. 

Coal and Ash Handling System at Pierce-Arrow Plant: Power, Jan. 13, 1914, p. 41. 

Coal and Ash Handling System at Victor Plant: Prac. Engr,, June 15, 1916. 

Mechanical Handling of Coal and Ashes in the Power Plant: Eng. Mag., Sept., 
1915, p. 872; Oct., 1915, p. 65. 

PROBLEMS. 

1. If power costs 1.5 cents per kw-hr. approximate the cost of moving 200 tons 
of coal per hour a horizontal distance of 50 ft. by means of a screw conveyor. 

2. Determine the power required to drive a scraper conveyor carrying 250 tons 
of bituminous coal per hour, sUding blocks to be used. The weight of the chain 
and flights with shding blocks is 26 lb. per Uneal ft., the capacity of the conveyor 
is 150 tons per hour. The distance between centers of head and last sprockets is 
160 ft. and the angle of conveyor with the horizontal is 30 degrees. 

3. Determine the power required to drive a pivoted bucket carrier having a ca- 
pacity of 60 tons of coal per hour; rollers 6 in. in diameter with If in. pins; weight 
per ft. of empty carrier, 80 lb.; horizontal length of conveyor, 400 ft.; vertical lift, 
60 ft.; 4 right angle turns; horizontal length traversed by loaded buckets, 300 ft.; 
speed of conveyor, 50 ft. per min. 

4. Determine the power required to elevate 200 tons of coal per hour by means 
of a 24-inch belt. Speed of belt, 200 ft. per min.; vertical lift, 30 ft.; length of con- 
veyor between centers, 300 ft. The system contains 3 fixed and 2 movable trippers. 



CHAPTER VII 

CHIMNEYS * 

126. General. — In order to cause the necessary weight of air to flow 
through the fuel bed and force the products of combustion through the 
gas passages of the boiler and setting, a pressure difference between 
ash pit and uptake is necessary. This pressure difference is designated 
as draft whether the actual pressure is above or below atmosphere. 
Draft may be produced mechanically by means of fans, blowers and 
steam jets, or thermally by means of chimneys. Stacks or chimneys 
offer the simplest means of conducting the products of combustion to 
waste and since the latter must be discharged at a sufficient elevation 
to prevent their being a public nuisance the height of stack necessary 
to effect this result is often sufficient to create the required draft. Even 
if considerable height must be added to the stack over and above that 
required to discharge the gases at a given elevation the extra cost may 
be considerably less than incident to mechanical draft operation. For 
this reason the majority of steam power plants depend upon the chim- 
ney for draft. In large plants equipped with mechanical stokers or 
where fuel is burned at a high rate or where economizers are used for 
abstracting heat from the flue gases mechanical draft is commonly 
employed; but even in these cases if forced draft is used some chimney 
effect may be desirable. In view of the enormous amount of heat 
developed in forced draft, stoker-fired furnaces and the great weight 
of gas passing over the boiler heating surfaces it is now generally ac- 
cepted that some means must be provided to remove these gases from 
the furnace promptly in order to protect the furnace brickwork, by 
preventing a '^ soaking up" action of the heat. The chimney pro- 
vides such a suction draft throughout all parts of setting. (See Para- 
graph 153.) 

When in operation, a chimney is filled with a column of gases with 
higher average temperature than that of the surrounding air. As a 
result the density of the gases within the stack is less than that of the 
outer air, and the pressure at the bottom of the column is less inside the 
stack than it is outside. 

* In this text the terms ''chimney" and "stack" are used synonymously. Build- 
ers usually apply the term "chimney" to the masonry and concrete structures and 
"stack" to the steel structures. 

279 



280 



STEAM POWER PLANT ENGINEERING 



TABLE 5L 



] 


DENSITY AND SPECIFIC VOLUME 


OF AIR AND CHIMNEY GASES AT 








VARIOUS TEMPERATURES. 








Air. 


Chimney Gases. 


t 


« 


V 


d 


t 


d 


t 


d 


t 


d 





11.581 


.935 


.086353 


200 


.06334 


430 


.04695 


660 


.03730 


5 


11.706 


.945 


.085424 


210 


.06239 


440 


.04643 


670 


.03697 


10 


11.832 


.955 


.084513 


220 


.06147 


450 


.04592 


680 


.03665 


15 


11.931 


.965 


.083623 


230 


.06058 


460 


.04542 


690 


.03633 


20 


12.085 


.976 


.082750 


240 


.05971 


470 


.04493 


700 


.03602 


25 


12.211 


.986 


.081895 


250 


.05887 


480 


.04445 


710 


.03571 


30 


12.337 


.996 


.081058 


260 


.05805 


490 


.04398 


720 


.03540 


32 


12.387 


1.000 


.080728 


270 


.05726 


500 


.04353 


730 


.03511 


35 


12.463 


1.006 


.080238 


280 


.05648 


510 


.04308 


740 


.03481 


40 


12.589 


1.016 


.079434 


290 


.05573 


520 


.04264 


750 


.03453 


45 


12.715 


1.026 


.078646 


300 


.05499 


530 


.04221 


760 


.03424 


50 


12.841 


1.037 


.077874 


310 


.05428 


540 


.04178 


770 


.03396 


55 


12.967 


1.047 


.077117 


320 


.05358 


550 


.04137 


780 


.03369 


60 


13.093 


1.057 


.076374 


330 


.05290 


560 


.04096 


790 


.03342 


62 


13.144 


1.061 


.076081 


340 


.05224 


570 


.04056 


800 


.03316 


65 


13.220 


1.067 


.075645 


350 


.05159 


580 


.04017 


900 


.03072 


70 


13.346 


1.077 


.074930 


360 


.05096 


590 


.03979 


1000 


.02861 


75 


13.472 


1.087 


.074229 


370 


.05035 


600 


.03942 


1100 


.02678 


80 


13.598 


1.098 


.073541 


380 


.04975 


610 


.03905 


1200 


.02516 


85 


13.724 


1.108 


.072865 


390 


.04916 


620 


.03869 


1300 


.02373 


90 


13.851 


1.118 


.072201 


400 


.04859 


630 


.03833 


1400 


.02245 


95 


13.976 


1.128 


.071550 


410 


.04803 


640 


.03798 


1500 


.02131 


100 


14.102 


1.138 


.070910 


420 


.04749 


650 


.03764 


1800 


.01848 


110 


14.354 


1.159 


.069665 










2000 


.01698 















d = density, pounds per cubic foot. 

I = temperature, deg. fahr. 

s = specific volume, cubic feet per pound. 

V = comparative volume, volume at 32 deg. fahr. = 1. 

Density of chimney gas taken 0.085 pound per cubic foot at 32 deg. fahr and 29.92 inches of mercury. 

(Rankine, " Steam Engine," gives the density at 32 deg. fahr. as varying from 0.084 to 0,087.) 

127. Chimney Draft. — The theoretical maximum static draft of a 
chimney is the difference in weight of the column of heated gas inside 
the stack and of a column of outside air of the same height, thus, if 
D = maximum theoretical static draft, in. of water, 
H = effective height of the chimney, ft., 
da = density of the outside air, lb. per cu. ft., 
dc = density of the inside gas, lb. per cu. ft., 
0.192 = factor for converting pressure in lb. per sq. ft. to in. of water, 

D = 0.192 H{ da - dc). (63) 

Neglecting the influence of the relative humidity of the air 

Pa T 



da = 0.0807 



P Ta 



(64) 



I 



CHIMNEYS 



281 



in which 

Pa = observed atmospheric pressure, lb. per sq. in., 
P — standard atmospheric pressure, lb. per sq. in., 
T = absolute temperature at the freezing point, deg. fahr., 
Ta = absolute temperature of the outside air, deg. fahr. 
The density of chimney gas varies with the nature of the fuel and the 
air excess used in burning the fuel. An average value is 0.085 lb. per 
cu. ft. at 32 deg. fahr. and pressure P. 

Therefore, p m 

dr = 0.0S5y'Yj (65) 

in which 

Tc = absolute temperature of the chimney gas, deg. fahr. 
Other notations as in equation (64). 
Substituting these values of da and dc in equation (63), 

n n 1 oo r/ ^« / 0.Q807 T 0.085 T \ ,_, 



TABLE 52. 



THEORETICAL 


DRAFT 


PRESSURE 


IN INCHES OF WATER. CHIMNEY 








100 FEET HIGH.i 










Temp. 


Temperature of the External Air — 


-Barometer, 14.7 Pounds per Square Inch.' 


in the 






















Chim- 
























ney. 


0^ 


lO*' 


20^* 


30« 


40° 


SO'' 


60° 


70° 


80° 


90° 


100° 


200 


.453 


.419 


.384 


.353 


.321 


.292 


.263 


.234 


.209 


.182 


.157 


220 


.488 


.453 


.419 


.388 


.355 


.326 


.298 


.269 


.244 


.217 


.192 


240 


.520 


.488 


.451 


.421 


.388 


.359 


.330 


.301 


.276 


.250 


.225 


260 


.555 


.528 


.484 


.453 


.420 


.392 


.363 


.334 


.309 


.282 


.257 


280 


.584 


.549 


.515 


.482 


.451 


.422 


.394 


.365 


.340 


.313 


.288 


300 


.611 


.576 


.541 


.511 


.478 


.449 


.420 


.392 


.367 


.340 


.315 


320 


.637 


.603 


.568 


.538 


.505 


.476 


.447 


.419 


.394 


.367 


.342 


340 


.662 


.638 


.593 


.563 


.530 


.501 


.472 


.443 


.419 


.392 


.367 


360 


.687 


.653 


.618 


.588 


.555 


.526 


.497 


.468 


.444 


.417 


.392 


380 


.710 


.676 


.641 


.611 


.578 


.549 


.520 


.492 


.467 


.440 


.415 


400 


.732 


.697 


.662 


.632 


.598 


.570 


.541 


.513 


.488 


.461 


.436 


420 


.753 


.718 


.684 


.653 


.620 


.591 


.563 


.534 


.509 


.482 


.457 


440 


.774 


.739 


.705 


.674 


.641 


.612 


.584 


.555 


.530 


.503 


.478 


460 


.793 


.758 


.724 


.694 


.660 


.632 


.603 


.574 


.549 


.522 


.497 


480 


.810 


.776 


.741 


.710 


.678 


.649 


.620 


.591 


.566 


.540 


.515 


500 


.829 


.791 


.760 


.730 


.697 


.669 


.639 


.610 


.586 


.559 


.534 


550 


.863 


.828 


.795 


.762 


.731 


.700 


.671 


.644 


.618 


.593 


.585 


600 


.908 


.873 


.839 


.807 


.776 


.746 


.717 


.690 


.663 


.638 


.613 



1. For any other height multiply the tabular figure by rrr-, where H is the height in feet. 

p 

2. For any other pressure multiply the tabular figure by , , — . where P is the barometric pres- 

14./ 
sure in pounds per square inch. 



282 STEAM POWER PLANT ENGINEERING 

Assuming Pa = P = 14.7 and T = 492, equation (66) reduces to 
^ „ /7.64 7.95\ ,^^, 

By assuming the same density for the chimney gas and outside air, 
and P = 14.7, equation (66) may be written 



D = 0.52 P, 



.(i--^)i?. (68) 



Equation (66) gives the true maximum theoretical static draft pro- 
vided the various factors entering into the formula are accurately known. 
In practice considerable variation exists in the composition of the 
gases and the temperatures are not uniform throughout the stack nor 
is the pressure the same at all points, hence the so-called theoretical 
value is correct only for an arbitrarily fixed set of conditions. Further- 
more, with a quiet atmosphere the theoretical draft may be largely in- 
creased owing to the column of heated gases above the mouth of the 
chimney. Strong air currents passing over the mouth of the stack 
may also increase or decrease the draft. The actual maximum static 
draft can be reahzed only when there is no flow as when the ash pit 
doors are closed and there is no perceptible transfer of heat or leakage 
of air through the chimney walls, boiler setting and flue or breeching. 

Example 16. Required the maximum theoretical draft obtained from 
a chimney 150 feet high, atmospheric pressure 14.5 pounds per square 
inch, temperature of outside air 60 deg. fahr., mean temperature of 
the chimney gases 550 deg. fahr. 

Here P„ = 14.5, Ta = 460 + 60 = 520, Tc = 460 + 550 = 1010, T = 
460 + 32 = 492. 

Substituting these values in equation (66) 

n A mo 14.5 /0.0807 X 492 0.085 X 492\ , ._ 
D = 0.192 j^(^ — j^j^— j 150 

= 0.994, or practically one inch of water. 

In this problem the mean temperature of the chimney gases is given. 
In practice it must be approximated from the flue gas temperature. 
Sufficient data are not available for predetermining the cooling action 
of the chimney walls and breeching except for a few special cases.* 
In view of the great variation in chimneys as to design, size, material, 
temperature difference and rate of driving, all assumptions are largely 
a matter of guess-work and equations based on a few isolated cases are 
equally untrustworthy. A common rule is to allow a drop of 80 de- 
grees per 100 feet of unlined steel stacks and 40 degrees for brick or 

* Peabody and Miller, Steam Boilers, p. 199. 



CHIMNEYS 



283 



lined steel chimneys. These values are too high for tall chimneys of 
large diameter and too low for small short stacks. Another rule is to 
allow 5 to 10 per cent of the theoretical maximum static draft as the 
pressure drop due to the cooling action. Both rules are purely arbitrary 
and may give results far from the truth. 

For the influence of rate of driving on stack temperatures see Table 
36 and Fig. 148. 

With economizers stack temperatures are reduced to 250-350 deg. 
fahr. Because of the increased height of stack necessary to neutralize 

650 



^575 



'550 



gC25 

a 

s 

£500 



475 



425 



























y^ 






A 11 High, 3 Pass, Vertical Baffle 
B 12 High. 3 Pass, Vertical Baffle 

High Stirling 

Rust 

12 High, 3 Pass, Horizontal Baffl 
C 9 High, 3 Pass, Vertical Baffle 

Low Stirling 

2 Pass Horizontal Baffle 

Wickes 












y 


y 












^ 






^ 








^^ 


^ 






^ 
















^ 


k> 


y^ 

B^ 


^ 


^ 


y^ 


^ 


^ 










^ 






^ 


^ 


C 


^ 












^ 








U^^ 


^ 


















J 






^ 






















■^ 























































100 



150 175 200 

Per Cent of Boiler Rating 



250 



Fig. 148. Relation between Flue Gas Temperature and Increase in Boiler Rating. 

Natural Draft. 

the reduction in stack temperatures economizer installations are com- 
monly made with forced or induced draft. 

As soon as a flow is established the static draft will decrease since 
part of this potential energy is required to impart velocity to the gases 
and overcome the resistance of the chimney walls. Furthermore, the 
breeching, boiler damper, baffles and tubes, and the bed and grate all 
retard the passage of the gases and the draft from the chimney is re- 
quired to overcome these resistances. If an economizer is used this 
adds a further pressure drop. (See paragraph 285.) Neglecting leak- 
age and minor influences, the various pressure losses may be expressed: 

/Z) = D, + D6 + Z)„ + D, + /^/ + D. + Dr, (69) 

in which / is an empirical coefficient depending largely on the rate of 
cooHng gases within the chimney, D is the maximum theoretical static 



284 



STEAM POWER PLANT ENGINEERING 



draft, Dg the pressure drop through the fuel and grate necessary to 
effect the desired rate of combustion, Db the drop through the boiler, 
Dv the draft required to impart velocity up to the damper, and Dd, D/, 
Dc, Dr, the respective draft losses through the damper, flue, chimney, 
and right angle turns into the breeching. Transposing equation (69) we 

have 2)^ _^ 2)^ + jT)^ + 2)^ = j2) - Dc - D/ - Dr. (70) 

Dg -{- Db -{- Dv T Dd is the draft required at the stack side of the damper. 
fD — Dc is the effective draft of the chimney and fD — Dc — Df — Dr 
is the available draft at the stack side of the damper. 

All these losses increase approximately with the square of the velocity 
of flow and may be expressed mathematically, but owing to extreme 
diversity in operating conditions many of the factors entering into the 
analysis can only be approximated, with the ultimate result that the 
calculated values are more or less arbitrary. Considering the losses in 
the order given in equation (69) : 

D, the total or maximum static draft, may be calculated from equa- 
tion (66). The limitations of this formula have been previously shown. 

Dg, the draft required to effect a given rate of combustion, de- 
pends upon the kind and condition of fuel, the thickness of fire, type 
of grate and eflBciency of combustion and can only be found accurately 
by experiment. For every kind of fuel and rate of combustion there 
is a certain draft with which the best general results are obtained. 



51. 

























































^ 


ii 




% 


ij 






/ 


























I 












•^j 


/ 


























t 






<^/ 






^ 


























^5 


?/ 




1 






/ 




























-V 






/7 






/ 




























/ 








/ 










yp 


f 
















1 


/ 




/ 




/ 


/ 











^ 


..'^f 
















/ 




/ 




/ 


/ 








6^ 


r 














/ 


/ 


/ 


/ 


/ 


/ 






■» 
■?^ 


?1y 


y$\ 




^^^t^ 


r^ 














/ 


/ 


/ 


y 


/ 






p 


^ 




"A^ 


^ 


" 














/ 


/ 


/ 


/ 


/ 






^ 


^ 


^ 


r 


















/ 




/ 


y 


^ 


^ 


^^ 




r 






















y 


'y 






^ 


r 


^■^ 


























^ 


^ 





































Q •" 
O .2 

I.. 



5 10 15 20 25 30 35 40 45 

founds of Coal burned per Square Foot of Grate Surface per Hour 

Fig. 149. Draft Required at Different Combustion Rates for Various Fuels. 



CHIMNEYS 



285 



The curves in Fig. (149)* give the furnace draft necessary to burn 
various kinds of coal at the indicated rate of combustion for average 
operating conditions. These curves allow a safe margin for economi- 
cally burning coals of the kinds noted. Specific figures for various 
types of stokers may be obtained from the manufacturers. 

TABLE 53. 

AVERAGE PRESSURE DROP THROUGH BOILER AiND SETTING. 
(Boilers Operating at 100 to 150 Per Cent Rating.) 



Boiler. 



Atlas, horizontal 2 pass, standard setting 

Atlas, vertical 3 pass, standard setting 

Babcock and Wilcox, Sewall baffle 

Babcock and Wilcox, vertical 3 pass, standard setting. 

Cahall, standard baffling and setting 

Continental, Dutch oven 

Edge Moor, vertical 4 pass, underground breeching 

Erie City, vertical boiler, standard baffling 

Hawkes, horizontal baffles, standard setting 

Keeler, vertical 3 pass, tube spacing 6 X 6 to 6 X 7. . . . 
Keeler, vertical 3 pass, tube spacing 5| X 7| to 5| X 6| 

Return tubular, old style setting 

Return tubular, double arch bridge wall setting 

Return tubular, McGinnis arches, front and back 

Stirling, standard setting and baffles 

Stirling, 5 pass 

Stirling, standard baffles, underground breeching 

Wickes, standard setting and baffles 

Worthington, standard setting and baffles 



25 
40 
35 
45 
35 
35 
36 
40 
40 
50 
45 
44 
53 
40 
35 
19 
14 
42 
45 



K = Percentage of the effective draft at the stack side of damper available in the combustion cham- 
ber. This factor applies only to hand-fired furnaces burning 20 to 30 pounds of Illinois coal per square 
foot of grate surface per hour and to mechanical stokers of the natural draft type burning 20 to 40 pounds 

u x?(t .■ A u Draft over fire X 100 
per hour. Effective draft = • 



Db, the loss of draft through the boiler and setting, varies within 
wide limits depending upon the type and size of boiler, arrangement 
of tubes and baffles, design of setting, type of grate, and rate of driving, 
and ranges from less than 0.1 inch to 1.0 inch and over. The data 
given in Table 53 t are based upon the investigations of 0. Monnett, 
former Chief Smoke Inspector of the City of Chicago, and may be used 
as a guide in predetermining the extent of these losses for different 
types of boilers and settings. The figures in the table apply to hand- 
fired grates having an air space of 45 to 55 per cent and rates of com- 
bustion ranging from 20 to 30 pounds of Illinois coal per square foot 
of grate surface. They also apply to mechanical stokers of the natu- 



* "Steam"; Babcock & Wilcox Co. 
t Power, June 2, 1914, p. 768. 



p. 246. 



286 



STEAM POWER PLANT ENGINEERING 



ral draft type, burning 20 to 40 pounds of coal per square foot of grate 
surface with the capacities in either case ranging from rating to 50 
per cent overload. The relative pressure drop increases with the 
load but there appears to be no close relationship between those two 



1.2 



1.0 



o 

O 0.6 

3 
w 

S 0.4 

PM 



0.2 



0.0 

































A 14 High, 3 Pass, Vertical Baffle 
B 12 High, 3 Pass, Vertical Baffle 

High Stirling 

Rust 

12 High, 3 Pass, Horizontal Baffle 
C 9 High, 3 Pass, Vertical Baffle 

Low Stirling 

2 Pass Horizontal Baffle 

Wickes 














/ 
















/ 


/ 
















/ 






















^ 


^ 




^ 


















A 


^ 






















^ 




^ 


■^ 








^ 










^ 


^-^ 


^ 




C^^ 


-^ 


-^ 
















— ' 





















































150 175 200 

Per Cent of Boiler Rating 



225 



Fig. 150. Pressure Drop through Boilers — Furnace to Stack Damper — 

Natural Draft. 



factors for different boiler equipments. Specific figures may be ob- 
tained from boiler manufacturers. The curves in Fig. 151 are plotted 
from a series of tests conducted by A. P. Kratz, on a Babcock & Wilcox 
boiler located in the new power plant of the University of Illinois. 
(A Study of Boiler Losses. A. P. Kratz, Bui. 32, University of IlKnois, 
April 12, 1915.) The futihty of assuming an ''average value for gen- 
eral practice" is evidenced from the extreme range even in this par- 
ticular installation. 

T>v, the draft required to accelerate the gases, varies in accordance 

with the law 

y,2 _ y 2 
^ = -^' (71) 

in which 



}i = head in feet of gas producing the velocity, 
V2 = initial velocity, ft. per sec, 
V^ = final velocity, ft. per sec, 

g = acceleration of gravity = 32.2 (approx.). 



CHIMNEYS 



287 



1.2 



cjc = Percentage of builders 

rating developed. 
" =Thicliness of fire, in. 




Fig. 151. Pressure Drop through Boiler Setting. 508 Horsepower B. & W. 
Boiler, 3-Pass Vertical with Tile-roof Furnace. — Chain Grate. 

Assuming a gas density of 0.085 lb. per cubic foot at 32 deg. fahr. 
and 14.7 lb. per sq. in. pressure, and reducing head in feet of gas to 
pressure in inches of water, equation (71) reduces to 



D. = 0.124 



Pa/T 



(E-a) 



(72) 



in which 

Fa = observed barometric pressure, lb. per sq. in., 
P = one standard atmosphere = 14.7 lb. per sq. in., 
Tc = absolute temperature of the chimney gases, deg. fahr. 

The draft required to accelerate at sea level from zero velocity to 
velocities 10, 20, 30, and 40 feet per second at a temperature of 550 



288 STEAM POWER PLANT ENGINEERING 

deg. fahr. is 0.012 in., 0.048 in., 0.108 in., and 0.192 in., respectively. 
Except at high velocities the draft is small and may be neglected in 
the problem of chimney design. 

Dd, the loss of draft through the damper, is varied arbitrarily to 
meet the load requirements. The minimum value of Dd corresponding 
to "wide open damper" is usually included in the boiler loss Di,. For 
the influence of damper area on the draft in fire-tube boilers, see "Draft 
in Fire Tube Boilers," S. H. Viall, Power, April 11, 1916, p. 509. See 
also, "Dampers for Water-Tube Boilers," Osborn Monnett, Power, 
May 26, 1914, p. 729. 

The commonly accepted rules for determining the friction loss Dc 
through the chimney are all based on Chezy's formula 



in which 



/P--' (73) 

2g m ^ 



f = coefficient of friction, 

I = length of the conduit, ft., 
m = mean hydraulic radius, ft., 
V = mean velocity of the gases, ft. per sec. 
Other notations as in equation (71). 

Because of the great variation in the assumed value of the coeffi- 
cient of friction, the rules referred to give widely discordant results 
for the same set of conditions. Satisfactory results have been ob- 
tained by assuming / = 0.012 for unlined steel stacks and / = 0.016 
for brick and brick-lined steel stacks. Until further experiments prove 
to the contrary these values may be accepted as being as accurate as 
any used in this connection. 

For circular chimneys, and square chimneys, D ft. square, Chezy's 
formula may be reduced to the convenient form 

D. = K y, (74) 

in which 

Dc = friction loss in inches of water, 

K = coefficient including the coefficient of friction and the various 
reduction factors, 
= 0.006 for unlined steel stack, 0.008 for brick or lined steel stack, 
V = velocity of the gases, ft. per sec, 
H = height of stack above the breeching, ft., 
D = diameter of the stack, ft., 
Tc = absolute temperature of the chimney gases. 



CHIMNEYS 289 

Considering the weight instead of the velocity of flow, equation (74) 
reduces to the form 

D. = k ^^-^, (75) 

in which 

/? is a coefficient including the coefficient of friction and the various 

reduction constants, 
W = weight of gas flowing, lb. per sec, 
d = diameter of the stack, inches. 

For the assumed values of / (0.012 and 0.016) : 
A; = 1.6 for unhned steel stacks, 1.9 for brick or brick-lined stacks. 

C. R. Weymouth, Trans. A.S.M.E., Vol. 34, 1912, gives k a value 
of 2.3 for fined and unlined stacks. 

According to Kingsley's experiments * the loss of draft due to skin 
friction, displacement of the atmosphere by the issuing stream and 
change of direction of the gases upon entering the stack conforms 
approximately to the following equation 

Dc = 0.00036 V\ (76) 

Notations as in equation (74). 

Equations (74) and (75) may be used for determining Df, the loss 
in the flue or breeching, by substituting the length of the flue for H 
and the diameter for d. A common allowance for the friction drop 
in flues, round or square, is 0.1 inch of water per 100 feet of straight 
conduit. 

Dr, the draft resistance due to right-angle turns, is ordinarily taken 
as 0.05 inch of water per turn. Another rule is to assume this resist- 
ance to be equivalent to a length of flue twelve diameters in length. 

An examination of equations (73) to (76) will show that the friction 
draft loss of the chimney cannot be calculated directly unless the 
height and diameter and weight or velocity of flow are known. Since 
these are the quantities to be determined it is evident that the prob- 
lem lends itself only to be a ''cut and trial analysis," provided the 
equations are to be satisfied. If the various pressure drops influencing 
the height of the stack could be calculated or estimated \vith any degree 
of accuracy there would be some reason for exact analysis, but the 
arbitrary values assigned in practice vary so widely that such analyses 
are ordinarily without purpose. Furthermore, the friction loss through 
the chimney is only a comparatively small percentage of the total 
loss (except for high velocities), hence a careful calculation of the chim- 

* Engineering Record, Dec. 21, 1907, p. 679. 



290 STEAM POWER PLANT ENGINEERING 

ney friction, and guess-work in estimating the other losses is highly 
inconsistent. Scattering tests made on a number of tall chimneys 
in successful operation show that the effective pressure at 100 to 150 
per cent rating is not far from 80 per cent of the theoretical maximum 
static pressure. Assuming this to hold true for chimneys in general, 
the problem of determining the height becomes a comparatively simple 
one. In view of the uncertainty of the coefficient of friction, results 
based upon this assumption are perhaps fully as rehable as those cal- 
culated from the various formulas. 

Example 17. Determine the height of a stack suitable for burning 
30 pounds of Illinois bituminous coal per sq. ft. of grate surface per 
hour for a hand-fired, return tubular boiler with double-arch bridge 
wall furnace, when the temperature of the outside air is 60 deg. fahr., 
the mean temperature of the flue gases 550 deg. fahr., and the flue 
is 100 feet long with two right-angle bends. 

This loss will be approximately as follows : 

Loss through fuel and grate (from curves in Fig. 149) 0.33 

33 
Loss in boiler (from Table 53), tt-To 0-62 

Loss in flue, 100 ft. at 0.10 per 100 0. 10 

Lossin turns, 2 X 0.05 0.10 

Total required draft at the breeching entrance to the stack 1.15 

On the assumption that the effective or required draft is 80 per cent 
of the theoretical maximum static draft. 

And from equation (67) 

from which 

H = 210 feet, height above damper. 

The vertical passes in any boiler act as chimneys and are capable 
of furnishing a draft pressure in much the same manner as the chimney 
proper. The greater the length of the vertical passes the greater will 
be the ''chimney action." The pressure difference due to the chimney 
action may decrease or increase the draft of the stack, depending upon 
the direction of flow of the gases. If the flow is upward the vertical 
pass acts as an additional height of stack, if downward, it tends to 
retard the flow. Thus, in the Wickes boiler, Fig. 56, the vertical path 
of the gases through the boiler itself causes considerable chimney 
action. At low rating the pressure at C may be atmospheric or even 
slightly above, although the draft in the combustion chamber B may 
be 0.10 inch of water below that of the atmosphere. This means 
that the boiler itself furnishes sufficient chimney action to operate the 



CHIMNEYS 291 

boiler at this load. Similarly the draft at D may be higher than at 
C due to the negative chimney action and resistance combined. The 
difference in temperature of the gases due to the cooUng action of the 
heating surface must of course be considered in calculating the chimney 
action. In practically all boilers the chimney action of the vertical 
passes influences the pressure drop throughout the setting and the 
effect is more marked when the rate of flow is low. See ''Draft in 
Furnaces and Flues," E. G. Bailey, Power, Nov. 9, 1915, p. 638. 

A well-designed central chimney serving several boilers and subject 
to considerable load variation should have comparatively low stack 
and breeching friction in order to insure ''draft regulation." While 
a certain draft margin is necessary it should be the aim to provide a 
chimney with the least possible excess draft over the necessaiy maxi- 
mum. For very high stacks, such as are required in tall office build- 
ings, the diameter is made very small so that a considerable portion 
of the pressure drop will occur in the stack and breeching, otherwise 
the draft will be excessive even with throttled damper. In designing 
stacks for this purpose the assumed draft loss in the stack and breech- 
ing should be made to conform with the law expressed in equation (75). 

Boiler Draft: Power, Mar. 20, 1917, p. 374; April 11, 1916, p. 509; Nov. 9, 1915, 
p. 638; Aug. 10, 1915, p. 196; May 18, 1915, p. 675; Jan. 12, 1915, p. 39; July 7, 
1914, p. 7; June 9, 1914, p. 806. 

The Significance of Drafts in Steam Boiler Practice: Bulletin 21; U. S. Bureau of 
Mines, 1911. 

Proportioning Chimneys on a Gas Basis: A. L. Menzin, Jour. A.S.M.E., Jan. 
1916, p. 31. 

Dimensions of Boiler Chimneys for Crude Oil: C. R. Weymouth, Trans. A.S.M.E., 
Vol. 34, 1912. 

Calculating the Dimensions of Chimneys and Stacks: G. A. Orrok, Power, Aug. 22, 
1916, p. 274; Sept. 12, 1916, p. 384. 

128. Chimney Area. — A study of equation (75) will show that any 
required effective draft may be obtained from various combinations of 
heights and diameters. Evidently there must be a certain height and 
diameter which will produce the cheapest structure. In practice this 
particular combination cannot be predetermined with an}^ degree of 
accuracy because of the uncertainty of the various factors entering 
into the problem of calculating the height and diameters. For an 
assumed set of conditions the logical procedure is to calculate a trial 
height for the required maximum rate of combustion, and then to 
proportion the area according to equation (75) so that the maximum 
weight of gases generated may be discharged at a rate corresponding 
to the assumed friction loss through the stack. By cut and trial a 
number of combinations of heights and diameters may be calculated 



292 STEAM POWER PLANT ENGINEERING 

in this manner which will give the required effective draft. The costs 
of the various structures may then be estimated and a selection made. 

In general practice this degree of refinement is seldom attempted 
and the usual procedure is to calculate a height compatible with the 
assumed pressure losses (subject, of course, to community laws) and 
proportion the area by rules which are more or less empirical. Thus, 
if the area is to be proportioned on a gas basis the maximum volume 
of the gases to be discharged is computed and an arbitrary velocity 
is assumed. 

If specific data are not available for computing the volume of the 
gases the area may be calculated by one of the various empirical equa- 
tions outlined in Table 56. 

Example 18: Proportion a brick-lined stack for water-tube boilers 
(vertical three-pass standard baffling) rated at 6000 horsepower, 
equipped with chain grates and burning Illinois coal; boilers rated 
at 10 square feet of heating surface per horsepower; ratio of heating 
surface to grate surface, 50 to 1; flue 100 feet long with two right- 
angle bends; stack to be able to carry 50 per cent overload; atmos- 
pheric temperature 60 deg. fahr.; sea level; temperature of flue gases 
at overload 540 deg. fahr.; calorific value of the coal 11,200 B.t.u. 
per pound. 

A modern plant of this type and size should be able to maintain a 
combined boiler, furnace and grate efficiency of 75 per cent at 150 
per cent rating. To be on the safe side assume it to be 70 per cent, 
then 

Maximum boiler horsepower = 6000 X 1.5 = 9000. 

Heat equivalent of 1 boiler horsepower-hr. = 34.5x970 = 33,479 B.t.u. 

33 479 
Coal per boiler hp-hr. = ^^ 2OO X 70 ^ ^'^ ^^' ^PP^^^' 

Total grate surface = zr^ = 1200 sq. ft. 

50 



38 700 
Maximum rate of combustion = ' ' = 32.3 lb. per sq. ft. grate 



Total coal burned per hour = 4.3 X 9000 = 38,700 lb. 

2.3 lb. pe 

surface per hour. 
Assumed pressure losses at maximum rating: 

Inches 
of Water. 

Loss through fuel and grate (from curves in Fig. 149) 0.34 

Loss in boiler (furnace to stack side of damper) . 55 

Loss in flue 100 ft. at 0.1 in. per 100 0. 10 

Loss in turns, 2 X 0.05 0. 10 

Total loss or required efTective pressure measured at flue en- 

trance of stack 1 . 09 

1 09 
Theoretical draft = -^-^ = 1.36 in. 



CHIMNEYS 



293 



Height of stack above damper, equation (67), 

/7.64 7.95\ 
^•^^ - 1520 ~ lOOOr • 

H = 202 ft. 

For 70 per cent combined efficiency the air excess with chain grate 
and Illinois coal may range from 50 to 75 per cent. To take care of 
possible reduction in efficiency, leakage and other adverse influences 
assume a total air excess of 100 per cent. 

Theoretical air per 10,000 B.t.u. = 7.5 lb. (See Table 13.) 



Theoretical air per lb. of coal = 7.5 



11,200 
10,000 



8.4 lb. 



Actual air per lb. of coal = 8.4 X 2 = 16.8 lb. 

Probable weight of flue gas per lb. of coal = 17.5 lb. 

(If the ultimate analysis of the coal is known the weight of the prod- 
ucts of combustion may be calculated as shown in paragraph (22). 
If the per cent of CO2 is assured this quantity may be calculated or it 
may be taken directly from Table 54.) 

w ' u, f » 17.5 X 38,700 ,QQ , 

Weight of flue gas = „^^„ = 188 pounds per sec. 



Total volume of flue gas 



3600 
188 
0.0418 



4500 cu. ft. per sec. 



(The density of the flue gas varies considerably with the nature of 
the fuel and the air excess.) 

TABLE 54. 
WEIGHT OF GASES FOR DIFFERENT PERCENTAGE OF CO2 WHEN CO = O. 



Per cent CO2 in the dry gases by vol- 
ume 

Excess air in per cent of the theoreti- 
cal minimum 

Weight of gases per 10,000 B.t.u 

Per cent of CO2 in the dry gases by 
volume 

Excess air in per cent of the theoreti- 
cal minimum 

Weight of gases per 10,000 B.t.u. in 
the coal 



18.7 



7.8 



18.0 

4.0 
8.1 

11.0 

68.0 

12.9 



17.0 

10.0 

8.6 

10.0 
85.0 
14.2 



16.0 

17.0 
9.1 

9.0 

105.0 

15.7 



15.0 

24.0 
9.6 

8.0 

130.0 

17.6 



14.0 

33.0 
10.3 

7.0 

162.0 

20.0 



13.0 

43 
11 

6.0 

206.0 

23.3 



12.0 

54.0 
11.9 

5.0 

267.0 

27.8 



TABLE 55. 

AVERAGE VELOCITY OF CHIMNEY GASES. 



Volume of chimney gases discharged, 
cu. ft. per sec 

Average velocity at maximum load, ft. 
per sec ' 



10 


100 


500 


2500 


5000 


8000 


10 


15 


20 


25 


30 


35 



12,000 
40 



These values are based upon data compiled from 200 modern chimney installations of various heights 
and diameters. There appeared to be no definite relationship between volume and velocity and the 
values in the table represent gross averages only. 



294' STEAM POWER PLANT ENGINEERING 

Assume 30 ft. per sec. as the average velocity of the gases. (See 
Table 55.) 

/irea = -^^ = 150 sq. ft. 

Corresponding diameter = 13.8 ft. or 165 inches. 

It will be noted that numerous assumptions have been made in the 
foregoing analysis, consequently the reUabihty of the results depends 
entirely upon the accuracy of these assumptions. Because of the 
possible variation in practice of these assumed values, and because in 
many situations they cannot be approximated with any degree of ac- 
curacy, many engineers prefer to proportion the area on such empirical 
equations as (5) and (12), Table 56. 

Thus, Kent's rule, equation (5), gives 

i?ff ,■ 0.3 X 9000 ^^ ^ ^^ _^ ,. 
Effective area = y= — X 0.86 = 163 sq. ft. 

V202 

Corresponding actual diameter = 177 inches. 

Kent's equation is based on a coal consumption of 5 lb. per boiler 
horsepower-hour, therefore 4.3 -^ 5.0 = 0.86 is the correction factor 
for the given conditions, hence the effective area as calculated from 
Kent's equation should be multiplied by 0.86. 

According to equation (12), Table 56, 

D = 4.92 hp.o-4 
= 4.92 X 60000-4 
= 160 inches. 

For small hand-fired plants it is sufficiently accurate to adopt the 
following proportions : 

Internal area of the chimney, one-fifth to one-sixth of the connected 
grate area for bituminous coal and one-seventh of the grate area for 
anthracite. 

The following heights have been found to give good results in plants 
of moderate size: 

Feet 

With free-burning bituminous coal 90 

With anthracite, medium and large sizes 120 

With slow burning bituminous 140 

With anthracite pea 150 

With anthracite buckwheat 175 

With anthracite slack 200 

For plants of 800 horsepower or more the height of stack for coal 
burning should never be less than 150 feet, regardless of the kind of 
coal used. Natural draft greater than 1.5. in. of water is seldom 
necessary and higher intensities can be obtained much better by forced 
or induced draft. This limits the height of chimney to about 225 
to 250 ft. 

In proportioning the area of the stack on a gas basis the data in Tables 



CHIMNEYS 295 

54 and 55 may be used as a guide. By plotting the data compiled from 
a number of modern chimneys the relation between velocity and area 
appeared to be approximately as follows: 

V = (0.2 + 0.005 D) V, (77) 

in which 

V = average actual maximum velocity of the chimney gases, ft. per 
sec, 

D = diameter of the chimney, ft., 

y = theoretical velocity, ft. per sec, assuming that the total theo- 
retical draft is available for producing velocity. 

139. Empirical Cliimney Equations. — The various empirical formulas 
outlined in Table 56 are occasionally used in proportioning chimneys. 
They give good results within the limits of the assumptions upon which 
they are based, but otherwise may lead to absurd results, their applica- 
bility depending largely upon the available data covering the various 
losses with the particular kind, quality, and condition of coal, and con- 
ditions of operation. Occasionally practical and local considerations 
fix the height of the stack irrespective of theoretical deductions. 

Referring to Table 56, equations (1), (2), (6), (7), and (9) are based 
upon a fuel consumption of 13 to 15 pounds of anthracite and 22 to 26 
pounds of bituminous coal per square foot of grate area per hour. In 
equations (3), (4), and (9), the diameter is dependent solely upon the 
quantity of coal burned per hour and the height is determined mainly 
by the rate of combustion per square foot of grate. The results accord 
well with practice. With western coals equation (3) gives results rather 
too large and the constant should be 120 instead of 180. Equation (5) 
is perhaps the most used and has met with much approval. It is based 
on the assumptions that: 

1. The draft of the chimney varies as the square root of the height. 

2. The retardation of the ascending gases by friction may be con- 
sidered due to a diminution of the area of the chimney or to a lining of 
the chimney by a layer of gas which has no velocity and the thickness 
of which is assumed to be 2 inches. Thus, for square chimneys, 

E = D^ -^ = A -IVJ, (78) 

and for round chimneys, 

E = l{^'-^) = ^- 0.591 Va. (79) 

For simplifying calculations the coefficient of VA may be taken as 0.6 
for both square and round chimneys, and the equation becomes 

^ = A - 0.6 VI. (80) 



296 STEAM POWER PLANT ENGINEERING 

3. The horsepower capacity varies as the effective area E. 

4. A chimney should be proportioned so as to be capable of giving 
sufficient draft to permit the boiler to develop much more than its 
rated power in case of emergencies or to permit the combustion of 
5 pounds of fuel per rated horsepower per hour. 

5. Since the power of the chimney varies directly as the effective 
area E and as the square root of the height H, the equation for horse- 
power for a given size of chimney will take the form 

Hp. = CE Vh, (81) 

in which C is a constant, found by Mr. Kent to be 3.33, obtained by 
plotting the results from numerous examples in practice. 
The equation then assumes the form 

Hp. = 3.33 E Vh, (82) 

or 

Hp. = 3.33 (A - 0.6 V J) Vh, (83) 

from which 

i/ = (°-^-)^ (84) 

Table 57 has been computed from equation 5, Table 56. 

130. Stacks for Oil Fuel. — In designing stacks for oil fuel or gas 
firing the procedure is the same as for coal burning, that is, the height 
is made sufficiently great to maintain the required draft in the furnace 
at maximum overload and the area is proportioned to take care of the 
maximum volume of gases generated. Excessive draft greatly influ- 
ences the economy of oil-fired furnaces, whereas with coal firing there 
is rarely danger of too much draft. Consequently greater care must 
be exercised in estimating the various draft losses through the boiler 
and breeching. With oil fuel there is practically no loss of draft through 
the fuel bed and grate and the pressure loss through the boiler will be 
less because of the smaller volume of gases discharged per boiler horse- 
power hour. Furthermore, the action of the burner itself acts to a 
certain degree as a forced draft. Therefore, both the height and area 
of the stack for a given capacity of boiler will be less for oil-firing than 
for coal-firing. Table 58 calculated by C. R. Weymouth (Trans. A.S. 
M.E., Vol. 34, 1912) after an exhaustive study of data pertaining to 
the subject may be used as a guide in proportioning stacks for oil fuel. 

131. Classification of Chimneys. — Chimneys may be grouped into 
three classes according to the material of construction: 

1. Steel. 

2. Reinforced concrete. 

3. Masonry. 



CHIMNEYS 



297 





















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298 



STEAM POWER PLANT ENGINEERING 



o 

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ry ^ >» -4, •— ' 
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p 

c a 

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CHIMNEYS 



299 



Steel chimneys have many advantages and are finding much favor 
in large power plants, especially where economy of space warrants 
the erection of the stack over the boiler, in which case the structural 
work of the boiler setting answers for both boiler and chimney. Among 
the advantages over the masonry construction are: (1) ease and rapid- 
ity of construction; (2) less weight for a given internal diameter and 
height; (3) less surface exposed to the wind; (4) lower cost; (5) smaller 
space required; (6) sHghtly higher efficiency if properly calked, for 
there can be no infiltration of cold air as is likely through the cracks 
in masonry. The chief disadvantage is the cost of keeping the stack 
well painted to prevent rust and the corrosive action of the sulphur 
in the coal. 

TABLE 58. 

STACK SIZES FOR OIL FUEL. 



Stack Diam- 
eter, Inches. 



33 
36 
39 
42 

48 

54 
60 
66 

72 
84 

96 
108 
120 



Height in Feet Above Boiler-room Floor. 



80 



161 

208 
251 
295 
399 

519 
657 
813 
980 
1373 

1833 
2367 
3060 



90 



206 
253 
303 
359 

486 

634 

800 

993 

1206 

1587 

2260 
2920 
3660 



100 



233 
295 
343 
403 
551 

720 

913 

1133 

1373 

1933 

2587 
3347 
4207 



120 


140 


270 


306 


331 


363 


399 


488 


474 


521 


645 


713 


847 


933 


1073 


1193 


1333 


1480 


1620 


1807 


2293 


2560 


3087 


3453 


4000 


4483 


5040 


5660 



160 



315 

387 
467 
557 
760 

1000 
1280 
1593 
1940 

2767 

3740 
4867 
6160 



Figures represent nominal rated horsepower; sizes as given are good for 50 per cent overloads, 
on centrally located stacks, short direct flues and ordinary operating efficiencies. 



Based 



Steel chimneys may be: 

1. Guyed. 

2. Self -sustained. 

13:?. Guyed Chimneys. — Guyed sheet-iron or steel chimneys or 
stacks held in position by guy wires are employed in small sizes on 
account of their relative cheapness. They seldom exceed 72 inches 
in diameter and 100 feet in height. A heavy foundation is unnecessary 
for the smaller sizes and the stack may be supported by the boiler 
breeching. The small short stacks are ordinarily riveted in the shop, 
ready for erection, larger sizes being shipped in sections and riveted 



300 



STEAM POWER PLANT ENGINEERING 



at the place of installation. In addition to a liberal allowance for cor- 
rosion the material is made heavy enough to support its own weight and 
to prevent buckling under initial tension of the guy wires and the stress 
due to wind action. The thickness of shell is ordinarily based on ar- 
bitrary rules of practice and no attempt is made to calculate this value 
by stress analysis. Table 59 gives the thickness of material as advo- 
cated by a number of manufacturers. 



TABLE 59. 
APPROXIMATE WEIGHT AND COST OF GUYED SHEET-STEEL CHIMNEYS. 



Height, Feet. 


Diameter, Inches. 


Thickness of Shell, 
B.W.G. 


Approxiniate Weight 
per Foot, Pounds. 


40 


18 


16 


13 


45 


20 


16 


14 


45 


22 


14, 16 


20, 15 


50 


24 


14, 16 


22, 16 


50 


26 


14 


23.5 


55 


28 


14 


25 


60 


30 


12, 14 


34,27 


65 


32 


12, 14 


36,28 


70 


34 


10, 12 


48,39 


75 


36 


10, 12 


51, 41 



Approximate cost per pound, 4 cents to 10 cents, including cost of sections 
riveted and punched, ready for assembling, the higher figure referring to the smaller 
stacks. 

Guy wires are furnished in one to three sets of three to six strands 
each and are attached to angle or tee iron bands at suitable points in 
the height of the stack. The lower ends of the guys are ordinarily 
anchored at angles of 50 or 60 degrees with the vertical. A rational 
analysis of the proper size of guy wires for a specified maximum wind 
pressure is impracticable because of the number of unknown variables 
entering into the problem, such as initial tension and stretch of the 
wires and flexure of the shaft. A common rule is to assume the entire 
overturning load to be resisted by one strand in each set of guys; thus, 
if there are two sets of guys the entire load is assumed to fall on two 
wires. An additional stress of one-half the overturning load is allowed 
for initial tension. A lattice bracing is frequently used between stacks 
when a number of stacks are placed in a continuous row. 

133. Self-sustaining Steel Chimneys. — Steel chimneys over 52 inches 
in diameter are usually self-supporting. They may be built with or 
without a brick lining, but the lining is preferred, since it prevents 
radiation and protects the inside from the corrosive action of the flue 



CHIMNEYS 



301 




"^Brick :^ackimj 



Floor L«vel 



302 STEAM POWER PLANT ENGINEERING 

gases. Since the lining plays no part in the strength of the chimney, it 
is made only thick enough to support its own weight, and usually of a 
low-grade fire brick or carefully burned common brick or both. In 
average practice the fire brick extends 20 or 30 feet above the breech- 
ing, the remainder of the hning being of common brick. In chimneys 
up to 80 inches internal diameter, the upper course is 4| inches thick 
and increases 4J inches in thickness for each 30 to 40 feet to the bottom. 
In larger chimneys about 8 inches is the minimum thickness. The 
lining is generally set in contact with the shell and thoroughly grouted, 
otherwise depreciation will be very great. 

In several recent designs vertical stiffeners are riveted to the shell 
which support horizontal rings or shelves on which the lining is built. 
The vertical stiffeners are spaced about 5 feet apart and the horizontal 
rings about 20 feet apart. By this method any section of the lining may 
be replaced without disturbing the rest. The lining is ordinarily of 
uniform thickness throughout the length of the shaft and seldom ex- 
ceeds 4 inches in thickness. 

Self-sustaining stacks may be straight or tapered, and are generally 
made with a flared or bell-shaped base whose diameter and length are 
IJ to 2 times the internal diameter of the stack. The base is riveted 
to a heavy cast-iron plate bolted to a concrete foundation of suffi- 
cient mass to insure stability. In the modern large station the stack is 
frequently carried on a steel structure over the boilers, thereby reducing 
ground space requirements. Such a design is illustrated in Fig. 130. 

Fig. 152 gives the details of one of the steel chimneys at the power 
house of the South Side Elevated Railroad, Chicago, Illinois. 

134. Wind Pressure. — Sufficient data are not available to show con- 
clusively the relation between wind velocity and the resulting effective 
pressure on surfaces of different shapes. Practically all authentic 
tests have been conducted on small flat surfaces and there is evidence 
to believe that the unit pressure exerted on large surfaces is somewhat 
less than that obtained from the former. Experiments conducted by 
different authorities show that the pressure per square foot of flat 
surface bears the following relationship to the wind pressure: 

in which P = KV^ 

K = coefficient determined by experiment, 
P = wind pressure, lb. per sq. ft., 
V = wind velocity, miles per hour. 

The value of K as determined by the different investigators varies 
from 0.0029 to 0.005. The most authentic tests give an average value 
K = 0.0032. This corresponds to a pressure of 32 lb. per sq. ft. of 



CHIMNEYS 303 

flat surface for a wind velocity of 100 miles per hour. Practically all 
chimneys are proportioned on a maximum wind velocity of 100 miles' 
per hour, but the unit pressure corresponding to this velocity is gener- 
ally assumed to be 50 lb. per sq. ft. of flat surface. Considering the 
unit pressure on a flat surface as 1, according to Rankine, the effective 
pressure for the same projected area is 0.75 for the hexagonal, 0.6 for 
octagonal, and 0.5 for round columns. Henry Adams, Industrial 
Engineering, 1912, p. 197, states that these figures are not in accord- 
ance with modern experiments and that the factors should be 0.785 for 
round and 0.82 for octagonal shafts. Current practice allows 25 to 
30 lb. per sq. ft. of projected area as the maximum unit pressure on 
round shafts. That 25 lb. per sq. ft. allows sufficient margin for safety 
is evidenced by the fact that chimneys proportioned on this basis are 
successfully withstanding the most \dolent gales. 

135. Thickness of Plates for Self-sustaining Steel Stacks. — If there 
is no wind blowing the only stress to be considered in the shell at any 
section is that due to the weight of the material itself, thus : 

S, = W^l{d,'-d,'), (85) 

in which ^ 

Si = stress (compression) due to the weight of the material, lb. per 
in. If the shaft is in perfect alignment this stress is uni- 
formly distributed over the entire cross section under con- 
sideration. 

W = weight of the shaft above the section under consideration, lb. 
If the fining is independent of the steel structure then the 
weight of the latter only is to be considered, but if the lining 
is supported by ledges secured to the shaft then the weight 
of the fining must be added to that of the steel. 

di = external diameter of the tube, in., 

c?2 = internal diameter of the tube, in. 

When the wind is blowing there is an additional stress due to bending. 

This is a tension on the windward side and a compression on the leeward 

side, thus, / 

S, = Ph^-, (86) 

in which ^ 

*S2 = stress in the outer fiber due to wind pressure, lb. per sq. in., 
P = the total wind pressure, lb., 
h = distance from the section under consideration to the center of 

wind pressure, in. For a cyfindrical shaft, h = i height of 

shaft above section. 



- = sectional modulus = -p^ { — — -. — - ] 
e 32 V rfi / 



304 STEAM POWER PLANT ENGINEERING 

The net stress, S, is therefore 

W Ph 

Equation (87) may be written 

^ _ [WW + d2') -^ 8] ± P/^ .^„. 

32 V d, I 

(di^ + di^) -^ 8 is commonly called the radius of the statical moment 
(see paragraph 143). Designating this "quantity by q, equation (88) 
reduces the convenient form 

S = {Wq ±Ph)^-' (89) 

Because of the liberal factor allowed for the safe working stress and 
because a tube of large diameter with thin walls will probably fail by 
flattening or buckling on the leeward side and not by tension of the 
windward side, the influence of the weight of the material is ordi- 
narily neglected and the shaft is treated as a cantilever subject to wind 
pressure only. Wq therefore is neglected and equation (88) becomes 

S = Ph-i-^-' (90) 

Since the thickness of the wall is a small fraction of the diameter 
the section modulus - becomes, approximately, 

- = 0.7854 d,% 
e 

in which 
t = thickness of the shell in inches. 
Substituting this value in equation (90) 

A number of steel stack builders simplify equation (91) still further 
by making the constant 0.8, thus 

Considering the stress, >S', per Hneal inch instead of that per sq. in. 
equation (92) becomes 



CHIMNEYS 



305 



Example 19: Determine the thickness of plate at a section 150 feet 
from the top of a cyUndrical steel stack 12 feet in diameter and 200 
feet high. Horizontal seams to be single riveted. 

The total wind pressure on the section is 

P = 150 X 12 X 25 * = 45,000 lb. 
The moment arm is 

h= "^ X 12 = 900 inches. 

S = 8000 lb. per sq. in. (A common allowance for safe stress is 
8000 lb. per sq. in. for single riveted and 10,000 for double 
riveted joints.) 
Substituting these values in equation (93) 

OQOQ - 45,000 X 900 
^^^^- 0.8 X 144^ ' 
from which 

t = 0.305. 

The nearest commercial size lies between ^^^ and VV- 

TABLE 60. 

STEEL STACKS. — SIZES OF RITER CONLEY COMPANY, PITTSBURG. 



Diameter 
of Flue. 


Total 
Height. 


Total 
Weight. 


Ft. In. 
5 6 


Ft. 
165 


Lb. 
67,000 


7 


160 


79,000 


8 6 
10 
12 


150 
200 
200 


94,000 
150,000 
175,000 


11 6 


225 


232,000 


12 


255 


256,000 



How Made. 



40 ft. of 



45 ft. of i in., 50 ft. of ^ in., 30 ft. of 



30 ft. of 1^ in., 50 ft. of i in., 50 ft. of i^ in., 30 ft. o: 

I in. 
60 ft. of i in., 60 ft. of ^ in., 30 ft. of f in. 
90 ft. of i in., 60 ft. of ^ in., 50 ft. of f in. 
35 ft. of \ in., 35 ft. of ^ in., 3.3 ft. of ^ in., 35 ft. 

of H in., 35 ft. of t in., 25 ft. of ^ in. 
40 ft. of i in., 40 ft. of ^ in., 40 ft. of ^ in., 40 ft. of 

H in., 40 ft. of I in., 25 ft. of ^ in. 
75 ft. of \ in., 65 ft. of ^ in., 55 ft. of | in., 35 ft. 

of ^ in., 25 ft. of ^ in. 



136. Riveting. — The diameter of rivets should always be greater 
than the thickness of the plate but never less than one-half inch. The 
pitch should be approximately 2^ times the diameter of the rivet, and 
always less than 16 times the thickness of the plate. Single-riveted 
joints are ordinarily used on all sections except the base, where the 
joint should be double riveted with rivets staggered, although in very 
large stacks all horizontal seams are double riveted to give greater 
stiffness to the shaft. 

* See Paragraph 133. 



306 



STEAM POWER PLANT ENGINEERING 



137. Stability of Steel Stacks. — For stability the resisting moment 
Wtq' must be greater than the Phi overturning moment (see paragraph 
143), that is 

Wtq' > Ph* (94) 

in which 

Wt = total weight of the structure, including that of the foundation, 
and the earth filling over the base, lb., 
q' = radius of the statical moment of the foundation base, ft., 
hi = distance from the center of wind pressure to the base, ft. 

For a square base the minimum value of qi, see equation (106), para- 
graph 143, is 

, L 



q = 



and the condition for stability is 



Wt^> Phi. 



(95) 



Expressed graphically: Lay off GP, Fig. 153, equal to the total wind 
pressure in direction and amount and acting at the center of pressure 
of the shaft; lay off GW equal to the weight of the stack and founda- 
tion; find the resultant GR and produce it to 
intersect the base line as at R'; if R^ falls within 
the inner third of the base the stack is stable, 
provided, of course, that the chimney is properly 
designed and constructed. Therefore the heavier 
the combined weight of the chimney and its 
foundation the more stable the structure. 

L in Fig. 153 varies from one-tenth to one- 
fifteenth H, depending upon the character of the 
subsoil. For the ordinary concrete foundation, 
Christie (''Chimney Design and Theory," p. 57) 
gives as an average value for L, 

+ 10. (96) 




Fig. 153. 



L = 



26,000 

138. Foundation Bolts for Steel Stacks. — There is no generally ac- 
cepted rule for proportioning foundation bolts for steel stacks. The 
various rules differ principally in the assumed location of the center 
of moments or neutral axis of the bolts when stressed by the over- 
turning moment. In lieu of proof to the contrary and considering 
the number of unknown factors entering into the problem the neutral 
axis may be taken as passing through and tangent to the bolt circle, 
* Axis of the shaft assumed to be vertical. 



I 



CHIMNEYS 



307 



and the fiber stresses in the bolts may be assumed to be proportional 
to their distances from the axis. Thus 

Ph-Wq = SaL, (97) 

in which 

Ph = wind moment at the base ring, in-lb., 
Wq = statical moment, in-lb., 
S = maximum fiber stress in the bolts, lb. per sq. in. (To allow 
for initial stress due to tightening up, a low fiber stress of 
12,000 lb. per sq. 
in. is commonly as- 
sumed.) 
a = area of each bolt at 
the root of the 
thread, sq. in. (All 
bolts assumed to 
be of the same 
diameter.) 
L = equivalent mean 
length of the bolt 
resisting moment, 
in. 
Referring to Fig. 154, 
SaL = S,h + 2 Soc + 2 S,d, (98) 
in which 




Fig. 154. 



Si, S2, Sz = stresses in bolts. A, B-B, and C-C, respectively, lb., 
h, c, d = respective moment arms relative to neutral axis XX, in. 

Since the stress in each bolt is assumed to be directly proportional 



to its distance from the neutral axis, S2 = Si r and S:i 

b 






Sub- 



stituting these values in equation (98) and noting that Si = Sa, equation 
(98) reduces to 



L = ^ (62 + 2c2 + 2(i2) 



(99) 



The value of L becomes 



Number of bolts 
L = bx 



6 

,2.25 



8 
3.00 



10 

3.88 



12 

4.58 



16 
G.OO 



24 36 
8.90 12.40 



Example 20: Calculate the size of bolts necessary for a steel stack 
with conditions as follows: Overturning moment 2,750,000 in-lb., bolt 
circle diameter 82 in., 6 bolts, allowable stress 12,000 lb. per sq. in. 



308 



STEAAI POWER PLANT ENGIXEERIXG 



Here Ph - Wq = 2,750,000; S = 12,000; L = 2.25 X 82 = 184.5. 
Substituting these values in equation (97) 

2,750,000 = 12,000 X a X 184.5; 
a = 1.24 sq. in. 

Nearest commercial size corresponding to this area, 1§ in. diam. 

Foundation Bolts for Steel Chimneys: D. A. Hess, Power, Oct. 5, 1915. 

Design of Steel Stacks: Eng. & Contr., Nov. 22, 1916, p. 440; Oct. 25, 1916, 
p. 369. 

Design and Construction of a 400-ft. Steel Stack: Eng. & Contr., Aug. 25, 1915, p. 140. 

Reasons for Corrosion of Steel Smokestacks and Ways to Prevent it: Elec. Wld., 
Nov. 6, 1915, p. 1033. 

139. Brick Chimneys. — By far the greater number of power-plant 
chimneys are of brick construction and usually of circular section, 
though octagonal, hexagonal, and square sections are not uncommon. 
The round chimney requires the least weight for stability, and the 
others in the order mentioned. 

Brick chimneys may be divided into two general classes: 

1. Single shell, Fig. 158, and 

2. Double shell. Fig. 156. 

The double shell is the more common and consists of an outer shaft 
of brickwork and an inner core or lining extending part 
way or throughout the entire length of the shaft. 

The single shell is the general construction where 
carefully burned and selected brick not easily affected 
by the heat are used. As the inner core or Unmg is 
independent of the outer shell and has no part in the 
strength of the chimney, the rules for determining the 
thickness of the walls are practically the same for both 
H single and double shell. 

140. Tliickness of Walls. — The thickness of the wall 
should be such as to require minimum weight of mate- 
rial for the proper degree of stability, due consideration 
being paid to the practical requirements of construc- 
tion. The thickness does not vary uniformly, but 
decreases from bottom to top by a series of steps or 
courses as in Fig. 155. In general, the thickness at 
any section should be such that the resultant stress of 
wind and weight of shaft will not put the masonry in 
tension on the windward side or in excessive compres- 
sion on the leeward side. 

For circular chimneys using common red brick for the outer shell 




Fig. 155. 



CHIMNEYS 309 

the following approximate method gives results in conformity with 
average practice: 

^ = 4 + 0.05 d + 0.0005 H, (100) 

where 

t — thickness in inches of the upper course, neglecting ornamenta- 
tion, and should, of course, be made equal to the nearest 
dimension of the brick in use. Ordinary red bricks measure 
8i X 4 X 2. 
d = clear inside diameter at the top, inches, 
H = height of stack, inches. 

Beginning at the top with this thickness, add one-half brick, or 
4 inches, for each 25 or 30 feet from the top downwards, using a batter 
of 1 in 30 to 1 in 36. 

The minimum value of t for stacks built with inside scaffolding 
should be 7 inches for radial brick and 8J inches for common brick, 
as a thinner wall will not support the scaffold. Radial brick for chim- 
neys are made in several sizes, so that the thickness of the walls when 
they are used increases by about 2 inches at the offsets. 

For specially molded radial brick or for circular shells reinforced as 
in Fig. 156 the length of the different courses may be much less than 
stated above. The external form of the top is a matter of appearance, 
and may be designed to suit the taste, but should be protected by a 
cast-iron or tile cap and provided with lightning rods. Ladders for 
reaching the top of the chimney are generally located inside the brick 
stacks and outside the steel structures. 

Professor Lang's rule (Eng. Rec, July 20, 1901, p. 53) for determin- 
ing the length of the different courses is (Fig. 155) : 

h = C (20t + mi-\- 0.1056 G + 2.5 ^ + 656 tan a - 0.007 H 

- 0.453 p - 18.7\ (101) 

in which 

h = length of the course under consideration, 

C = constant = 1 for a circular, 0.97 for an octagonal, and 0.83 for 

a square, chimney, 
i = increase in thickness for each succeeding section in feet, 
G = weight per cubic foot of brickwork, 
p = wind pressure, pounds per square foot, 
a = angle of the internal batter. 

All other notations as indicated in Fig. 155. 

For chimneys over 100 feet in height he recommends that 100 be 



310 



STEAM POWER PLANT ENGINEERING 




Fig. 156. Brick Chimney at the Power Plant of the Armour Institute of Technology. 



CHIMNEYS 311 

used instead of the actual height, since the critical point will be in 
one of the lower sections and not at the base. 

If a value of h is obtained which is not contained an even number 
of times in H, it may be slightly increased or decreased so as to effect 
this result. 

To determine the stresses at any section the shaft is treated as a 
cantilever uniformly loaded with a maximum wind pressure of 25 
pounds per square foot. If the tension on the windward side sub- 
tracted from the compression leaves a positive remainder, the chimney 
will be stable; if the remainder is negative, the masonry will be in 
tension, which it withstands but feebly. The sum of the compressive 
stresses on the leeward side due to wind pressure and weight must be 
less than the crushing strength of the masonry. The practice, however, 
of assuming a fixed value for allowable pressure irrespective of the 
height of the stack gives dimensions that are too low for small stacks 
and too high for large stacks. According to Professor Lang, compres- 
sive stress on the leeward side in pounds per square inch with single 
chimneys should not exceed 

p = 71 + 0.65 L, (102) 

where 

p = pressure in pounds per square inch, 

L = distance in feet from top of chimney to the section in question. 

With double shell p = 85 + 0.65 L. (103) 

The tension on the windward side should not exceed, 

for single shell: p = (18.5 + 0.056 L), (104) 

for double shell: p = (21.3 + 0.056 L). (105) 

Example 21. Determine the maximum stress in the outer fiber of 
the brickwork at the base of section 8 of the chimney illustrated in 
Fig. 158 when the wind is blowing 100 miles an hour. Assume the 
weight of the brickwork 120 pounds per cubic foot. 

A wind velocity of 100 miles per hour is estimated to exert a pres- 
sure 25 pounds per square foot of projected area on a cylindrical surface. 
(See paragraph 133.) The height of the chimney to section 8 is 131.4 
feet. The projected area as computed from the figure is 1800 square 
feet. Hence p, the total wind pressure, is 1800 X 25 = 45,000 pounds. 
The volume of brickwork above section 9 may be calculated, and is 
6150 cubic feet, hence the weight W = 6150 X 120 = 738,000 pounds. 

The area of the joint at this section is 75.3 square feet, therefore the 
pressure due to the weight of the superimposed brickwork is 738,000 
divided by 75.3 = 9800 pounds per square foot. To find the stress 
due to the wind pressure, substitute the proper values in equation (86) : 



Ph = S-= 0.0983 
e 



r-^)- 



312 STEAM POWER PLANT ENGINEERING 

Here 

P = 45,000 as computed above, 

h = 55 feet (found by laying out the section and locating the center 

of gravity), 
di = 16,2, 

d = 12.9, 

whence 

1 A 94 _ 1 9 04 

45,000 X 55 = 0.0983 ^q 2 "^' 

from which S = 9907 pounds per square foot. 

The net stress on any part of the section is the resultant of that due 
to the weight of the stack and that caused by the wind, the net stress 
on the windward side being 

9907 - 9800 = 107 pounds per square foot, 

which is evidently a tensile stress and should never exceed the value 
given by formula (104) : 

p = (18.5 + 0.056 L) 
= (18.5 + 0.056 X 131.4) 
= 25.8 pounds per square inch 
= 3715 pounds per square foot. 

The net compressive stress on the leeward side is 9800 + 9907 = 
19,707 pounds per square foot, which should not exceed that given by 
formula (102) : 

p = 71+0.65L 
= 71+0.65 X 131.4 
= 156.4 pounds per square inch 
= 22,521 pounds per square foot. 

141. Core and liining. — The core or lining of a brick chimney is 
commonly carried to the top of the shaft, though it sometimes extends 
only part of the distance. The inside diameter is generally uniform, 
the offsets being made on the outside. The core and outer shell should 
be independent to prevent injury due to expansion of the core. The 
rules for the thickness of Hning in steel chimneys apply also to brick 
chimneys. The batters for the inner and outer shells should be such 
as to allow at least 2 inches clearance between the two shafts at the 
top, and the top should be protected by an iron ring or by a projecting 
ledge from the outer shell. 

143. Materials for Brick Cliimneys. — Brick for the external shaft 
should be hard burned, of high specific gravity, and laid with lime 
mortar strengthened with cement. Lime mortar itself is more resist- 
ant to heat, but hardens slowly and may cause distortion in newly 
erected tacks, and hence should be used only when a long time is 



CHIMNEYS 



313 



taken in building. Mortar of cement and sand alone is not to be recom- 
mended, since it does not resist heat well and is attacked by carbon 
dioxide, particularly in the presence of moisture. A mortar consisting 
of 1 part by volume of cement, 2 of lime, and 6 of sand may be used for 
the upper brickwork, 1, 2J, and 8 respectively for the lower part, and 1, 1, 
and 4 respectively for the cap. The harder the brick the more cement is 
necessary, as lime does not cling so well to hard, smooth surfaces. The 
inner core may be constructed of second-class fire brick, since the tem- 
perature seldom exceeds 600 deg. fahr. Lime mortar is invariably used 
for the core. 

143. Stability of Bricli Chimneys. — When there is no wind blowing 
and the chimney is built symmetrically about a vertical axis the pres- 
sure due to weight is uniformly distributed 
over the bearing surfaces, and the center of 
pressure lies in the line XX, Fig. 157. But 
when the wind blows the pressure exerted 
tends to tilt the shaft as a whole column in 
the direction of the current, and the re- 
sultant pressure at the windward side of the _ h 
base decreases, until, with a sufficiently high 
velocity of wind, it may become zero, in 
which case the center of pressure moves a 
distance q towards the leeward side of the 
base. As soon as the pressure at A be- I I ^ 
comes zero the joint begins to open (assum- 
ing no adhesion between chimney and base) 
and the shaft is evidently in the condition 
of least stabihty. The distance q through 
which the center of pressure has moved is called the radius of the statical 
moment. For any column it may be shown that 




Fig. 157. 



g = -7- (Rankine, '' Apphed Mechanics, " p. 229), (lOG) 

in which 

I = moment of inertia of the section, 
A = area of the section, 

e = distance from the center of the shaft to the outer edge of the 
joint. 

D 

s' 

L 



Thus for a soHd circular section, q = 



For a solid square section, 



« = 6- 



314 STEAM POWER PLANT ENGINEERING 

For an annular circular ring, q = ^ — 

For a hollow square, q = —^ — 

The relationship between weight of shaft and wind pressure for the 
condition of least stabihty is 

Ph = Wq, (107) 

in which 

P = total wind pressure, pounds,' 

h = distance in feet from the base Une of the section under con- 
sideration to center of gravity of that section, 
W = weight of shaft in pounds above the assumed base line, 
q = radius of the statical moment. 

The condition of least stabihty for round chimneys requires, there- 
fore, that 

Ph^w'^j^' (108) 

For many purposes it is sufficiently accurate to assume D = dy and 
equation (77) becomes 

Ph — W -r^ for round chimneys, (109) 

Ph = W — for square chimneys. (110) 

o 

Another rule gives for the condition of least stability: 

W {\R^-\r) =Ph. (Eng. Rec, July 27, 1901, p. 82.) (Ill) 

Notations as in Fig. 108, all dimensions in feet. 

This permits of a lighter chimney than equation (108), and the maxi- 
mum wind pressure may be assumed to put the joint on the windward 
side in tension or even to permit a slight opening of same. 

A rule of thumb for stability is to make the diameter of the base one- 
tenth of the height for a round chimney; for any other shape to make 
the diameter of the inscribed circle of the base one-tenth of the height. 

The factor of stabihty is the quotient obtained by dividing the value 
of q from formula (107) by that from (106). If less than unity, the 
chimney is in tension at the outer fiber on the windward side, and must 
be redesigned unless the tension is less than that allowed by equation 
(104). Calculations for stability should be made for various sections. 

Example 22. Analyze the chimney illustrated in Fig. 158 for stability 
at, say, section 8, the following data referring to the portion above the 
base line of this section. 



CHIMNEYS 



315 



U-ii-^^^ 



I'r-s^iq- 



3H-*l^ 




-2 m^ 



20J^^ 



S.= 



Top of 
Foundation ^ 



10 



i 



-B 



TOTAL HEIGHT 

ABOVE FOUNDATION 

200 FT. 




SECTIO 



SECTfON ON ArA 

Fig. 158. Custodis Radial Brick Chimney. 



316 STEAM POWER PLANT ENGINEERING 

From the drawing: 

Projected area of the stack, 1800 square feet. 
Volume of brickwork, 6150 cubic feet. 
Outside diameter of base, 16.2 feet. 
Inside diameter of base, 12.9 feet. 
Center of pressure to base hne, 55 feet. 
Total height above base line, 131.4 feet. 
Maximum total wind pressure: 

P = 1800 X 25 = 45,000 pounds. 
Weight of shaft: 

W = 6150 X 120 = 738,000 pounds. 
For stabihty, according to equation (55), 

D2 -I- r/2 

Substituting the proper values: 

Ph = 45,000 X ^55 = 2,475,000 foot-pounds. 






SD 



(1 A 92 _|_ 1 9 Q2\ 
8 X 16.2 ) = 2.441,000. 



/)2 _J_ (P 

While Ph is sHghtly greater than W — ^ , for practical purposes 

the shaft at this section would be called stable under maximum allow- 
able wind pressure. 

For stability, according to equation (HI), 

Ph<W(iR+ir), 

Ph = 2,475,000, as determined above, 

= 4,177,000. 

Ph is therefore considerably less than W ii R -\- ir), and the con- 
dition imposed in equation (111) is more than fulfilled. 

The Design of Tall Chimneys: Henry Adams, Industrial Engineering, March, 1912, 
p. 198. Design of a Brick Chimney: Eng. News, May 9, 1912, p. 866. 

144. Custodis Radial Brick Chimney. — Fig. 158 gives the details of 
a 200 X 10-foot radial brick chimney constructed of special molded 
radial brick, formed to suit the circular and radial lines of each section, 
thus permitting them to be laid with thin, even mortar joints. The 
blocks are much larger than common brick and the number of joints is 
proportionately reduced. They are molded with vertical perforations, 
as shown in Fig. 159, which permits thorough burning, thereby in- 
creasing the density and strength and at the same time reducing the 



CHIMNEYS 



317 





n 



weight of the block. In laying, the mortar is worked into the per- 
forations about one-half inch. The first 60 feet above the base are 
octagonal in section, with 36-inch walls, and the balance of circular 
section, with walls tapering gradually from 22 inches to 7| inches in 
thickness. A radial brick Uning extends 60 feet from the base as in- 
dicated. The chimney was designed 
to furnish draft for a 3500-horse- 
power boiler plant and cost, erected, 
$8,800. The entire weight of the 
chimney exclusive of foundation is 
870 tons. 

Radial brick chimneys without 
the inner lining are likely to be 
unduly affected by temperature 
changes. 

The largest chimney of this type is located at Great Falls, Mont., 
and is used for leading off the gases from the smelter plant of the Boston 
and Montana Consolidated Copper and Silver Mining Company. 
The height above the top of the foundation is 506 feet, and the internal 
diameter at the top 50 feet. The chimney and foundation cost approx- 
imately $200,000. ; , 




BBO 



Fig. 159. Custodis Radial Brick. 



Custodis Chimney Details: Eng. Rec. 
p. 12. 



Oct. 1, 1904, p. 385; Power, May, 1900, 



145. Wiederholt Chimney. — This type of chimney consists essen*- 
tially of a combination of the masonry and reinforced concrete struc- 
tures. The inner and outer surfaces of the shaft are formed by 

hard burned fire clay tile of special design as illus- 
trated in Fig. 160. When placed in position these 
tile form a permanent mold into which the reinforc- 
ing bars and concrete may be introduced. Both 
vertical and horizontal reinforcing bars are incor- 
porated in the structure in much the same manner 
as in the Weber type. Because of the tile Uning 
much higher temperatures may be safely carried 
than with concrete type and the color may be 
readily made to match that of the power house or adjoining buildings. 

146. Steel-Concrete Chimneys. — The use of concrete reinforced with 
iron or steel for the construction of chimneys is rapidly increasing. The 
advantages claimed for this class of stack are: 

1. Light weight of the whole structure, being but one-third as great 
as an equivalent common brick chimney. The space occupied is much 




Fig. 160. Tile for 
Wiederholt Chimney. 



318 



STEAM POWER PLANT ENGINEERING 



O 1 .— 



-.4-(a- - - 



o "^ 




Si 



%• 



DETAIL OF HEAD 




jj Horizontal rings @ 
14 centers nlong entire 
ight of shaft, wound 
-spirallj 



SECTION AT TOP 





i extra vertical bars 
each side of opening 



;i^ |) Horizontal rings® 
centers along entire 
height of shaft and 
I lining, wound spirally 



Space 52 - 54 (j) 
ITS @ S'/j" centers 



"All vertical bars to be 
placed at least 2 "from 
surface of concrete 



SECTION AT BASE OF SHAFT 
^-t 20-0-^^ >\ 





\ 




1^^ 


if^ 




fVy L. 


Aa^Yi 




/SM 


^V' 


^dOG 
















^^ 


^^ 


x^ 




>^«x 


svKSj 


i^^ 


^^^ 


z 








1^^ 



FOUNDATION REINFORCEMENT 
Rectangular net • 5< 'h @ 14'cti*. 
riagonal net .. « « 7* .i 



Fig. 161. Weber "Coniform" Reinforced Concrete Chimney. 



CHIMNEYS 319 

less than with either brick or steel stack, on account of the thinness of 
walls at the base and the absence of any flare or bell. 

2. Total absence of joints, the entire structure including foundation 
being a monolith. 

3. Great resisting power against tension and compression. 

4. Rapidity of construction. May be erected at an average rate of 
six feet per day. 

5. Adaptability of the material to any form. 

This type of chimney being comparatively new, little data concerning 
depreciation are available, but some which have been in use ten years 
show little or no deterioration. 

Fig. 161 gives the details of a Weber ''coniform" steel-concrete 
chimney as erected at Grafton, Mass., for the Grafton State Hospital. 
The entire structure, foundation, shaft and lining is monolithic, 157 
feet in total height, seven feet internal diameter and weighs only 344 
tons. It occupies but 108 square feet of ground space at grade level. 
The weight of the shaft and lining is 249 tons. 

The shaft is of the double shell type with inner core extending 65 feet 
above the grade. The core is but 4 inches in thickness and the shaft 
varies from 10y\ inches at the junction of the core and shaft to 4 inches 
at the top. The core reinforcement consists of twelve vertical §-inch 
twisted steel bars and similar horizontal bars wound spirally at 14-inch 
centers. The vertical reinforcement in the outer shell varies from fifty- 
two f-inch twisted bars at the grade to twelve f-inch bars at the top. 
The horizontal reinforcement consists of |-inch twisted steel rings 
spaced at 14-inch centers along the entire height of shaft and wound 
spirally. The steel bars vary from 16 to 30 feet in length and where 
they meet lengthwise are lapped not less than 24 inches. The use of 
different lengths of steel prevents the laps from concentrating in any 
given section. 

The tallest chimney in the world is of this type and is located in 
Japan. It is 567 feet high and 26 feet 3 inches in diameter at the top. 

The determination of the amount of steel reinforcement does not 
permit of simple mathematical calculation because of the number of 
variables entering into the problem and graphical charts plotted from 
semi-rational formulas offer a simple solution. The curves in Fig. 162 
are reproduced from ''Principles of Reinforced Concrete," p. 408, 
Turneaure and Maurer, and are used extensively in this connection. 
The use of the chart is best illustrated by a specific example. 



320 



STEAM POWER PLANT ENGINEERING 



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Fig. 162. Wind Stresses in Steel-Concrete Chimneys.. (Turneaure and 

Maurer.) 



CHIMNEYS 321 



Strain Sheet. 

Example 23. Determine the amount of reinforcement required for 
the chimney illustrated in Fig. 161 at section BB. 
From the drawing we find: 

D = 11 ft. 9| in. r = radius of the steel circle = 5.79 ft. 

d = 10 ft. H in. h = 153 ft. 

The following values may be obtained by simple arithmetic computa- 
tions, but the actual calculation will be omitted for the sake of brevity. 

W, weight of shaft above section BB, 409,000 lb. 

A, area of shaft above section BB, 4320 sq. in. 

M, wind moment above section BB, 2,600,000 ft-lb. 

M 
e, eccentricity = Tj^ = 6.36 ft. 

5=1.1. 

r 

Assume a maximum compression in the concrete of fc = 360 lb. 
per sq. in. (In practice this assumed value varies from 350 lb. per 
sq. in. for chimneys under 150 ft. in height to 500 lb. per sq. in. for 
chimneys 350 ft. high.) 

fcA 
m, a coefficient = -—- = 3.8. 

From the curves in Fig. 162 the intersection of m = 3.8 and - = 1.1 

r 

gives p (per cent of steel required) as 0.53. 

T^ ^ area steel 

But v = -; 

area section 

Whence area of steel = 0.0053 X 4320 = 23 sq. in. corresponding 
to 52, f-inch steel bars. 

Other sections at 20 ft. intervals have been analyzed in a similar 
manner and the results inserted in Fig. 161. 

In the earlier types of steel concrete chimneys designed and built 
by the Weber Company the amount of steel reinforcement was cal- 
culated from formula (89), but all recent structures are proportioned 
on the Turneaure and Maurer chart. The resultant stress R as cal- 
culated from equation (89) necessitates the use of more reinforcement 
than that derived from the chart. 

Evase Stacks. See paragraph 153. 

Design, Construction, and Cost of a lS7-ft. Reinforced Concrete Chimney: Eng, & 
Contr.,Aug. 11, 1915,p. 111. 

147. Breeching. — The area of the flue or breeching leading from the 
boilers to the chimney is generally made equal to or a little larger than 
the internal area of the chinmey at the top, 10 per cent greater being an 



322 



STEAM POWER PLANT ENGINEERING 



average figure. A common rule is to allow 1 to 4| or 5 as the ratio of 
breeching area to grate area for ordinary service, and 1 to 3| for large 
boilers operating continually at 150 to 200 per cent rating. The flue 
may be carried over the boilers or back of the setting or even under 
the fire-room floor, but in any case should be as short as possible and 
free from abrupt turns. Underground breechings cause excessive 
pressure drop and are difficult to clean. Short right-angled turns 
reduce the draft approximately 0.05 inch for each turn, and a convenient 
rule is to allow 0.1 inch loss for each 100 feet of flue if of circular cross 
section and constructed of steel, and double this amount for brick flues 
of square section. Each additional boiler connected to the breeching 



r^ 


^^ 

















i__r "I r ^ 



One Boiler Leading 
Off to Side 



Two or More Boilers Leading Off 
to Side 





Two or More Boilers with Stack 
in Center. 



Fig. 163. Types of Breeching Connections. 



will cause a pressure drop due to friction or interference of the gases as 
they enter the breeching, or to leakage through the dampers when the 
boiler is out of service. A common rule is to allow a pressure drop of 
0.05 in. of water for each boiler connected to the breeching. The cross 
section of the flue need not be the same throughout its entire length, 
but may be tapered and proportioned to the number of boilers. Where 
two flues enter the stack on opposite sides, a diaphragm is inserted as 
indicated in Fig. 158. Flues should be covered on the outside with 
heat-insulating material, because a lining on the inside is diflicult to 
repair and deterioration might readily escape detection. A damper 
ratio of 1 to 4 expressed in terms of grate surface has given good satis- 
faction. 

148. Chimney Foundations. — On account of the concentration of 
weight on a small area the foundation of a chimney should be carefully 
designed. In most cities the building laws Hmit the maximum loads 



CHIMNEYS 323 

allowed for various soils and materials, and although they vary con- 
siderably the average is approximately as follows: 

Material, Safe Load, Lb. per Sq. Ft. 

Hard-burned brick masonry, cement mortar, 1 to 2 20,000-30,000 

Hard-burned brick masonry, cement mortar, 1 to 4 18,000-24,000 

Hard-burned brick masonry, lime mortar 10,000-16,000 

Concrete, 1 to 8 8,000-10,000 

Kind of Soil. Safe Load, Tons per Sq. Ft. 

Quicksands and marshy soils 0.5 

Soft wet clay 1.0 

Clay and sand 15 feet or more in thickness 1.5 

Pure clay 15 feet or more in thickness 2.0 

Pure dry sand 15 feet or more in thickness 2.0 

Firm dry loam or clay 3 . 0-4 . 

Gravel well packed and confined 6 . 0-8 . 

Rock broken but well compacted 10 . 0-15 . 

SoUd bed rock Up to 5 of its ultimate crushing strength. 

Tons per Pile. 

Piles in made ground 2.0 

Piles driven to rock or hardpan 25 . 

Chimney foundations as a rule are constructed of concrete except 
where the low sustaining nature of the soil necessitates the use of piles 
or a grillage of timber or steel. For masonry chimneys the foundation 
is designed to give the necessary support to the shaft without particular 
reference to its mass or distribution, as the shape of the foundation has 
virtually no effect on its stability as a column. In steel and reinforced 
concrete chimneys the shape and weight of the foundation are a func- 
tion of the desired factor of stabihty, since the shaft is securely anchored 
to the foundation and the two form practically one mass. The founda- 
tion should be designed to fulfill the conditions for shear and flexure 
in addition to the requirements for stability. 

Practically all chimney foundations are square in plan and the 
maximum pressure on the supporting surface may be calculated from 
the following equation : * 

4 W 

3 (1 - t-') b^ 
in which ^ / 

P = maximum pressure due to wind and weight, lb. per sq. in., 

W = total weight of the chimney and foundation, lb., 

. M . 
e = eccentric ^^ = wind moment divided by the weight, 

h = width of the foundation. 

* Principles of Reinforced Concrete Construction, Turnoauro and Maurer, p. 423. 



324 



STEAM POWER PLANT ENGINEERING 



For the chimney illustrated in Fig. 158 

4 X 738,000 



P = 



3 1 



= 3900 lb. per sq. ft. 

7S8 000 
The pressure due to weight only is ' — = 1845 lb. per sq. ft. 

Table 61 gives the least diameter and depth of foundation for steel 
chimneys of various diameters and heights. 



TABLE 6L 

SIZES OF FOUNDATION FOR STEEL CHIMNEYS. 



Diameter, Feet. 


Height, Feet. 


Least Diameter of 
Foundation. 


Least Depth of 
Foundation. 


3 


100 


15' r 


6' 0' 


4 


100 


16' 4" 


6' 0" 


4 


125 


18' 5" 


7' 0" 


5 


150 


20' 4" 


9' O'' 


5 


200 


23' 8" 


10' 0'^ 


6 


150 


21' 10" 


8' 0'^ 


6 


200 


25' 0" 


10' O'' 


7 


150 


22' r 


9' 0'' 


7 


250 


29' 8" 


12' 0" 


9 


150 


23' 8" 


10' 0" 


9 


275 


33' 6" 


12' 0" 


11 


250 


24' 8^ 


10' 0* 


11 


350 


36' 0" 


14' 0" 



149. Chimney Efficiencies. — The chimney as a mover of air has a 
very low thermodynamic efficiency. Compared with that of a fan its 
performance is very poor, and mechanical-draft concerns sometimes 
use this as an argument. 

Example 24. A chimney 200 feet high and 10 feet in diameter fur- 
nishes draft for a battery of boilers rated at 3500 horsepower. Average 
outside temperature 60 deg. fahr.; temperature of flue gases 500 deg. 
fahr.; calorific value of the fuel 14,000 B.t.u. per pound. Compare the 
thermal efficiency of the chimney as a mover of air with that of a 
forced-draft apparatus of equivalent capacity. 

From Table 52 we find that a chimney 200 feet high, with tempera- 
tures as stated above, will furnish a theoretical draft of 1.27 inches, 
equivalent to a pressure of 6.6 pounds per square foot. Neglecting 
friction, the height H oi Si column of external air which would produce 
this pressure is 



H 



(t>. 



(113) 



CHIMNEYS 325 



in which 

h = height of the chimney in feet, 
d = density of the hot gases in the stack, 
di = density of the outside air. 

Substitute in equation (113) 

di = 0.0763, d = 0.0435, and h = 200. 



/ 0.0763 - 0.0435 \ 
^-[ 00763 )^^^^ 



= 85.9 feet. 

The theoretical velocity of the air entering the base of the chimney 

under this head is /- — j^ 

V = V2 gH 

= V2 X 32.2 X 85.9 
= 74.5 feet per second. 

The weight of the gas escaping per second 

= 74.5 X area of the stack X 0.0763 
= 446 pounds. 

The displacement of this volume of gas is the result of heating it 
from 60 to 500 deg. fahr. Taking the specific heat of the gas as 0.24, 
the heat necessary to displace 446 pounds per second is 

Heat required = 446 X 0.24 X (500 - 60) 
= 47,000 B.t.u. per second. 

The work actually performed is that of overcoming a total resistance 
of 6.6 X 78.5 = 518 pounds (78.5 = internal area of the chimney) 
through a space of 74.5 feet; i.e.. 

Work done = 74.5 X 518 = 38,591 foot-pounds per second 

= 49.7 B.t.u. per second. 

49.7 
Efl&ciency = . „ = 0.00107, or about j\ of 1 per cent. 

If a fan be substituted for the chimney and we allow say 8 per cent 
for the efficiency of engine and boiler, 40 per cent for the fan, and 25 
per cent for friction, the combined efficiency will be 

0.08 X 0.40 X 0.75 = 0.024, or 2.4 per cent. 

024 
The fan then will be /^p>ir>7 = 22.4 times more efficient than the 

chimney as a mover of air. 

150. Cost of Chimneys. — Christie ('' Chimney Design and Theory") 

gives the following costs of chimneys 150 feet high and 8 feet internal 

diameter: 

Common red brick approximate cost $8,500 . 00 

Radial brick do. do. 6,800 . 00 

Steel, self-supporting, full lined do. do. 8,300.00 

Steel, self-supporting, half hned do. do. 8,800.00 

Steel, self-supporting, unlined do. do. 5,820.00 

Steel, guyed do. do. 4,000.00 



326 



STEAM POWER PLANT ENGINEERING 



The following approximate costs of various sizes of a well-known 
radial brick chimney give an idea of the variation in cost due to in- 
crease in diameter and height: 

TABLE 62. 



Size of Chimney. 




Size of Chimney. 








Cost. 






Cost. 


Height. 


Diameter. 


Height. 


Diameter. 


Feet. 


Feet. 




Feet. 


Feet. 




75 


4 


$1,350.00 


175 


8 


$7,050.00 


75 


6 


1,950.00 


175 


10 


7,925.00 


75 


8 


2,650.00 


175 


12 


8,950.00 


75 


10 


3,725.00 


175 


14 


9,725.00 


125 


6 


3,500.00 


200 


8 


9,250.00 


125 


8 


4,250.00 


200 


10 


10,500.00 


125 


10 


4,675.00 


200 


12 


11,100.00 


125 


12 


5,125.00 


200 


14 


12,500.00 


150 


.8 


6,150.00 


250 


10 


16,500.00 


150 


10 


7,125.00 


250 


12 


18,250.00 


150 


12 


7,750.00 


250 


14 


21,500.00 


150 


14 


8,275.00 


250 


16 


24,250.00 



PROBLEMS. 

1. Determine the maximum theoretical draft obtainable from a chimney 200 ft. 
high; altitude 2250 ft. (barometer 27.5 in.); temperature outside air 80 deg. fahr.; 
temperature of the flue gas 500 deg. fahr. 

2. Calculate the height of stack suitable for burning 20 lb. of anthracite buck- 
wheat per sq. ft. of grate surface per hr. for a hand-fired return tubular boiler, stand- 
ard setting, when the temperature of the outside air is 70 deg. fahr. and that of the 
flue gas is 450 deg. fahr. Assume a pressure loss in the boiler of 0.45 in. 

3. Determine the height and diameter of stack for a battery of Wickes vertical 
water-tube boilers rated at 4000 horsepower, equipped with chain grates and burning 
Illinois screenings; boiler rated at 10 sq. ft. of heating surface per hp.; ratio of heat- 
ing surface to grate surface 65 to 1; flue 50 ft. long; stack to be able to carry 100 
per cent overload; atmosphere temperature 60 deg. fahr., average barometric pres- 
sure 29 in.; temperature of flue gas at overload 650 deg. fahr.; calorific value of 
the coal 11,000 B.t.u. per lb. Assume pressure drop through boiler from the curves 
in Fig. 150. 

4. Determine the thickness of plates at various sections for a self-supporting 
steel stack of the height and diameter as calculated in Problem 3. 

5. Determine the size of foundation for the chimney in Problem 4. 

6. Design a brick chimney suitable for the data in Problem 3. 
Analyze the various sections for strength and stability. 



CHAPTER VIII 

MECHANICAL DRAFT 

151. General. — The intensity of natural draft in a chimney depends 
mainly upon the height of the stack and the temperature of the chimney 
gases, and the chimney should be designed to meet the maximum 
requirements, permitting the damper to be partly shut at times. There 
is usually no practicable means of increasing natural draft per se after 
the maximum has been reached. Again, chimney draft is peculiarly 
susceptible to atmospheric influence and may be seriously impaired 
by adverse winds and air currents. Notwithstanding these apparent 
limitations, by far the greater number of steam power plants depend 
upon chimneys for draft because of the disposition of the waste gases. 
In many cases artificial draft has a great advantage and under certain 
conditions is indispensable; it is very flexible and readily adjusted to 
effect various rates of combustion, irrespective of climatic influences, 
and permits any degree of overload without undue expenditure of 
energy. 

Artificial draft may be broadly classified under three heads: 

1. The vacuum or induced draft. 

2. The plenum or forced draft, and 

3. The ''balanced" draft method. 

In the induced draft system a partial vacuum is produced above 
the fire by suitable apparatus, and the effect is substantially that of 
natural draft. 

In the forced-draft system pressure is produced in the ash pit, the 
air being forced through the fuel bed. 

The so-called ''balanced draft" system is a combination of forced 
draft and induced or chimney draft. The pressure created by forced 
draft is made sufficient to overcome the resistance of the fuel bed while 
the chimney or induced draft is depended upon for creating a suction 
throughout the furnace and setting. The adjustment is such that 
practically atmospheric or a shght suction pressure exists in the com- 
bustion chamber. 

In all these systems the artificial draft is usually produced by either: 

1. Steam jets, or 

2. Centrifugal fans or exhausters. 

327 



328 



STEAM POWER PLANT ENGINEERING 



153. Steam Jets. — Fig. 164 shows an application of a ring jet to 
the base of a stack. The apparatus is very simple, inexpensive in 
first cost, and easily apphed. It consists essentially of a ring or a 
series of concentric rings of 1-inch pipe, perforated on the upper side 





Fig. 164. Ring Steam Jet. 



Fig. 165. Bloomsburg Jet. 



with XF" ^^ 8-inch holes, and placed in the base of the stack, so that 
the jets are discharged upward, thus creating a draft independent of 
the temperature of the flue gases. The steam connection to the jet is 




Fig. 166. McClaves Argand Blower. 

generally made direct to the boiler and not to the steam main, though 
the jet is often produced by exhaust steam. 

Fig. 165 illustrates a Bloomsburg jet, which involves to some extent 
the principle of the ejector. 



MECHANICAL DRAFT 



329 



The increase in draft produced by these devices as ordinarily in- 
stalled is not great, although in locomotive practice where the entire 
exhaust is discharged up the stack an intense draft is obtained. 

Fig. 166 shows the apphcation of a ''McClaves argand blower." 
The steam is discharged below the grate through a perforated hollow 
ring, as indicated, drawing the air through the funnel by inspiration. 
This creates a powerful draft by forming an air pressure in the ash 
pit, and is an especially useful system of forcing fires for boilers which 
need forcing for short periods only. 

Steam jets, as ordinarily installed, are very uneconomical, since a 
large amount of steam is required to produce good results. Table 63, 
based on experiments at the New York Navy Yard, to determine the 

TABLE 63. 
RESULTS OF EXPERIMENTS UPON STEAM JETS AT NEW YORK NAVY YARD.* 





Pounds of Water Evaporated per Hour. 


Index of Jet. 


A 


B 


C 


D 


E 


In boiler making steam 

In boiler supplying jets 

Per cent of steam used 
bv iet 


463.8 
97.5 

21.2 


580.0 
120 

20.7 


361.25 
30 

8.3 


528.5 
63.2 

12.0 


545.00 
76.25 

19 







* Annual Report of the Chief of the Bureau of Steam Engineering, U. S. Navy, 1890. 



TABLE 64. 
CONSUMPTION OF STEAM BLASTS COMPARED, t 



Cxjal. 


Name of Blower. 


Per Cent of Air 

Openings in 

Grate. 


Pounds of Dry- 
Coal burned per 
Hour per Square 
Foot of Grate. 


Per Cent of Total 
Steam Generated 

in the Boilers 
that is required 

to operate the 
Steam Blasts. 


Rice 


Young 




25.8 
17.9 
27.0 
27.3 
16.7 
31.4 
16.4 
26.1 
32.5 
45.4 


11 1 


Do 

Do 


.. .do 

Wilkinson 

Young ,...'.. . 


7.0 
10 8 


Buckwheat 


10 8 


Do 


. . . .do 


4.6 


Do 


do 


8 9 


Do 


McClave 

...do 


6 7 


Do 


9.3 


Do 


Wilkinson 

do 


7.8 


Do 


10.2 









t Trans. A.S.M.E., Vol. XVII. — See Whitham. 



330 



STEAM POWER PLANT ENGINEERING 



best form of steam jet for producing draft in launch boilers, shows 
steam consumptions of from 8.3 to 21.2 per cent of the total steam 
made. Table 64 gives the steam consumption of a number of types 
of steam jet blowers as determined by A. J. Whitman. The best 
performance is 4.6 per cent and the poorest 11.1 per cent of the total 
boiler steam generated. Steam jets below the grate are said to prevent 
clinkers from forming where fine anthracite coals are used, and thus 
to assist in keeping the fire free and open. They also assist in the 
economical combustion of certain low-grade fuels. See paragraph 93 
for the influence of steam jets in effecting smokeless combustion. 





































































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20 40 60 80 100 120 140 160 180 200 

Dry Coal per Square Foot of Grate per Hour - Poimds 

Fig. 167. The Relation between Draft and Rate of Combustion. Consolidated 

Locomotive. 



The curves in Fig. 167 are of interest in showing the intensity of 
draft created by steam jets in the modern locomotive and the influence 
of the draft on the rate of combustion. These curves are taken from 
Bulletin 2, Univ. 111., Sept. 13, 1915, p. 16. 

In large modern central stations where boiler overloads of from 150 
to 250 per cent above rating are desirable, steam jets and mechanical 
blowing and stoking apphances use but a nominal percentage of the 
steam generated. The results in Table 65, taken from the tests of the 
large Stirling boilers at the Delray Station of the Detroit Edison Com- 
pany, show what may be expected from installations of this class 
(Jour. A.S.M.E., Nov., 1911). 

153. Fan Draft. — Fig. 168 shows a typical installation of a centrif- 
ugal fan on the forced-draft or ^plenum principle, the fan creating a 



MECHANICAL DRAFT 



331 



TABLE 65. 

STEAM CONSUMPTION OF DRAFT APPLIANCES AND STOKER ENGINES. 2365 H.P. 

STIRLING BOILER. 
(Delray Station, Detroit Edison Co.) 

RONEY STOKER. 



No. of 


Per Cent 
of Rating. 


Dry Coal 

per Sq. Ft. 

G. S. per 

Hr. 


Steam Consumption, Per Cent 
of Total Generated. 


Draft, Inches of Water. 


Test. 


Stoker 
Engines. 


steam 
Jets. 


Total. 


Below 
Dampers. 


In 
Furnace. 


Ash 

Pit. 


5 

4 

18 


94 
152 

195.7 


14.81 
25.97 

33.60 


0.19 

0.15 
0.13 


1.56 
1.43 
1.19 


1.75 
1.58 
1.32 


0.16 
0.55 
1.11 


0.24 
0.22 
0.33 


0.10 
0.02 
0.05 



TABLE Q5 — Concluded. 



TAYLOR STOKER. 





Per Cent 
of Rating. 


Dry Coal per 

Sq. Ft. G. S. 

per Hr. 


Steam 

Consumption 
of Stoker 

Engines and 
Turbine 
Blower. 


Draft, Inches of Wate.-. 


No. of 
Test. 


At Blast in 
Tuyeres. 


Suction Below 

Boiler 

Dampers. 


Suction in 
Ash Pit. 


10 

9 

11 


92.9 

162.8 
211.0 


16.43 

29.23 
38.75 


2.63 
2.87 
3.41 


0.67 
1.73 
2.53 


0.20 
0.53 
84 


0.15 
0.06 
0.02 



All of the steam exhausted from the Taylor equipment may be returned to the feed-water heater, 
whereas only that exhausted from the engines in the Roney equipment may be used in this manner, hence 
the net heat used is approximately the same in both cases. 

For application of steam jets to mechanical stokers see Chapter IV. 



pressure in the ash pit and forcing air through the fueL The most 
approved method is to pass the air through the bridge wall, thence 
toward the front of the grate, though it may enter through an under- 
ground duct or through the side of the setting. Forced draft is usually 
adopted in old plants where increased demands for power require that 
the boilers be forced far above their rating to save the heavy expense 
of new boilers, or in plants burning refuse, anthracite culm or screen- 
ings, which require an intense draft for efficient combustion. Forced 
draft is also well adapted for underfeed stokers of the retort type, 
hollow blast grates, and the closed fire-hole system. The air supply 
may be taken from an air chamber built around the breeching, thereby 
supplying the heated air to the fan and effecting a lower temperature 
in the breeching and a higher temperature in the furnace. The ob- 
jection is sometimes raised against forced draft that the gases tend to 



332 



STEAM POWER PLANT ENGINEERING 



pass outward through the fire door when the fire is cleaned or replen- 
ished, since the pressure in the furnace is greater than atmospherio. 
This objection may usually be overcome by suitable dampers in the 
blast pipe which are closed on opening the fire doors or by having 
sufficient stack action to create a partial vacuum in the combustion 
chamber. With a boiler plant of 1000 horsepower or more the cost 
of a forced-draft fan, engine, and stack will approximate from 20 to 
30 per cent of the outlay for an equivalent brick chimney. The power 
consumption will depend upon the character and efficiency of the motor 
or engine and will range from 1 to 5 per cent of the total capacity. 




Fig. 168. Typical Forced-draft System. 



Induced draft as illustrated in Fig. 168 is perhaps the most com- 
mon substitute for natural draft and is extensively used in street rail- 
way and fighting plants which have high peak loads, being ordinarily 
installed in connection with fuel economizers. The suction side of the 
fan is connected with the uptake or breeching of the boiler or batteries 
of boilers and the products of combustion are usually exhausted through 
a stub stack. The illustration shows a typical installation in which 
two fans of the duplex type are placed above the boiler setting. The 
fan ducts are generally designed with a by-pass direct to the stack to 
be used in case of accident or when mechanical draft is not required. 

Since the fan handles hot gases it must, under the ordinary con- 
ditions of practice, have a capacity approximately double that of a 
forced-draft fan defivering cold air, but the gases being of lower density 
the power required per cubic foot moved is less. 

With forced draft from 200 to 300 cubic feet of air are required per 



MECHANICAL DRAFT 



333 



pound of coal; with induced draft the fan must handle twice this volume 
if the gases are exhausted at 500 deg. fahr. or 300 to 450 cubic feet if 
exhausted at 300 deg. fahr., a temperature to be expected in connection 
with economizers. 

The advantages of induced draft over forced draft are very pro- 
nounced. The pressure in the furnace is less than atmospheric, there- 
fore it is not necessary to shut off the draft in cleaning fires of ash pit, 
and the fire burns more evenly over the entire grate area, since the 




Fig. 169. Typical Induced-draft System. 

draft pressures are ordinarily less than with forced draft. An induced- 
draft plant costs considerably more than forced draft on account of 
the larger fan required, but the operating expenses are but httle greater. 
With a boiler plant of 1000 horsepower or more the cost of a single 
induced-draft fan, engine, stack, etc., will approximate from 40 to 50 per 
cent of the outlay required for a brick chimney of equivalent capacity, 
and the double-fan outfit will approximate from 50 to 60 per cent. 
The double-fan system is particularly adapted to plants which operate 
continuously and where even a temporary break-down is a serious in- 
convenience. 



334 



STEAM POWER PLANT ENGINEERING 



[« 6'0 



Turbo-undergrate draft blowers, installed in each setting, are finding 
favor with many engineers because of the low cost of installation. 

They consist essentially of small impulse steam turbines direct con- 
nected to specially designed propeller fans set in the side walls of the 
setting by means of wall thimbles. The fan discharges below the 
grate, and may be automatically controlled by damper regulation. 
The turbine exhaust may be discharged into the ash pit to prevent 
clinkers, or it may be used in the feed-water or other heating devices. 
They are more economical in heat consumption than the ordinary jet 
device. 

In Europe induced draft created by a fan discharging into the base 
of an evase stack is finding favor with many engineers. A few instal- 
lations have been made in this country by the 
Schutte and Koerting Company, but data rela- 
tive to their performance are not available. In 
this system a short stack (seldom exceeding 70 ft. 
in height) and resembling a Venturi tube is fitted 
with a small pressure blower near the base. The 
stack action is based on the injector principle and 
is sufiicient to operate the boiler at low rating 
without the use of the fan. For higher ratings 
air is discharged into the stack just below the 
Venturi throat and the suction in the breeching 
is greatly increased. These stacks are usually 
applied to single boilers or batteries. The general 
dimensions of an evase stack as installed in the 
power plant of the IngersoU Rand Company, 
Philhpsburg, Pa., is shown in Fig. 170. The fan 
requires about 5 per cent of the rated boiler horse- 
power for operation and the static pressure of the 
blower is approximately eight times the draft re- 
quirements in the breeching. 



-Air Supply 
Control 



Fig. 170. Evase Stack 
Capable of Furnish- 
ing Draft for the 
Combustion of 6000 
lb. of Coal per hour 
with Maximum Suc- 
tion at Breeching of 
1.5 Inches of Water. 



Mechanical Draft and the Evase Sta^k: Eng. Mag., July, 
1915, p. 525. 



Tall chimneys are a necessity in most cities since legislation requires 
the gases to be discharged at a height above that of adjacent buildings. 
In such situations, with stokers of the forced-draft type, tall stacks or 
induced draft would at first thought appear to be a necessary evil. 
Experience, however, shows that suction draft is an important factor in 
effecting efficient combustion and in prolonging the life of the furnace 
brickwork. By mutually adjusting the pressure created by the forced- 



MECHANICAL DRAFT 



335 



araft apparatus and the suction of the chimney or its equivalent, a 
so-called '^balanced-draft^' effect can be produced in the combustion 
chamber; that is, the pressure in the combustion chamber becomes 
practically atmospheric. The relative pressure drops are shown graph- 
ically in Fig. 171. This condition of positive pressure under the fire 
bed, zero or slightly suction pressure in the combustion chamber and 
a suction pressure throughout the rest of the setting (1) prevents dis- 
charge of the furnace gases into boiler room through leaky fire doors, 
inspection doors and cracked settings; (2) minimizes stratification and 
short circuiting of the air supply and combustible gases; (3) reduces 
the ^'soaking up" action of heat by the furnace brickwork; (4) assists 
reduction of air excess and (5) effects increase in overall boiler, furnace 




213 



a d 






Atmospheric Pressure 



Fig. 171. Pressure Drops through Boiler — Combined Forced Dni."; ;::. 1 Chimney. 



and grate efficiency. Many of our modern central stations are oper- 
ating with practically balanced-draft conditions as will be seen from the 
data in Table 66. In these plants the stoker speed, fan speed and stack 
damper are automatically controlled so as to effect the desired result. 
In the Essex Power Station of the Pubhc Service Electric Company, 
New Jersey, which is representative of the very latest practice (1917) 
the chimneys are 250 feet high and are served with both forced- and in- 
duced-draft fans. The induced-draft fan gives a maximum suction in 
the uptake of two inches of water pressure and the forced-draft equip- 
ment is capable of maintaining a pressure of six inches water under 
the grates. After the gases have passed from the boiler this may be 
discharged directly into the stack or by closing proper dampers in the 
breeching can be made to pass through the economizer and then to the 
stack; by closing a second damper the gases will pass through the 



336 



STEAM POWER PLANT ENGINEERING 



induced-draft fans before going to the stack. This makes it possible 
to operate the boilers under the most economical conditions at all times. 




Air 
Duer» 



Fig. 



-Air Supply 'Damper • 

172. McLean ''Balanced-draft" System. 



The term ''balanced draft" as applied to furnace work originated 
with Embury McLean and refers strictly to his system of control, see 
Fig. 172, but the term is now applied to any system in which chimney 
and fan draft are controlled so that the pressure in the combustion 
chamber is approximately atmospheric. 




1 1 ] 



Fig. 173. Mechanical Draft as Applied to Waste-Heat Boiler for Open-Hearth 

Furnace. 



MECHANICAL DRAFT 



337 



TABLE 66. 

DRAFT PRESSURES IN MODERN CENTRAL STATION AT OVERLOADS. 
(Underfeed Stokers.) 



Plant. 


Type of 
Boiler. 


TjDe of 

Stoker. 


Boiler 

Rating, 

Per Cent 

of Rated 

Capacity. 


Height of 
Stack 
Above 
Breech- 
ing. 


Flue 
Tempera- 
ture, Deg. 

Fahr. 


Static Draft, Inches of 
Water. 


Ash 
Pit. 


Com- 
bustion 
Cham- 
ber. 


Stack 
Side of 
Dam- 
per. 


Delray No. 1 

Delray No. 2 

Delray No. 2 

Connors Creek 

Boston Elevated 

59th St. Interborough . . 
74th St. Interborough . . 
74th St. Interborough . . 


Stirling 
Stirling 
Stirling 
Stirling 
B. &W. 
B. &W. 
B. &W. 
B. & W. 


Taylor 
Taylor 
Taylor 
Taylor 
Taylor 
Taylor 
Riley 
Westing- 
house 


*175 
*175 
220 
*175 
240 
200 
335 
292 


242 
196 
196 
240 
165 
200 
242 
242 


550 
600 
620 
580 
515 
523 
631 
609 


+3 5 
+3.5 

+4.2 
+3.8 
+4.0 
+3.3 

+5.8 
+3.8 


-0.03 
-0.03 
-0.07 
-0.10 
-0.03 
-0.14 
-0.02 
-0.15 


-1.2 

-1.2 

-1.3 

-0.8 

-0.76 

-0.52 

-0.74 

-0.65 



* Normal operating maximum. 



Draft and Stoker Control at Waterside: Power, Nov. 7, 1914, p. 698. 
Boiler Control Boards at Delray: Power, Sept. 28, 1915, p. 435. 
The Essex Power Station: Power, Nov. 28, 1916, p. 739. 

Performance of Boilers with Balanced Draft: Elec. Wld., Sept. 9, 1916, p. 522; Aug. 
12, 1916, p. 321. 

154. Types of Fans. — Centrifugal fans for mechanical draft may- 
be divided into two general classes; those having rotors with a few 
straight or slightly curved blades of considerable length radially, Fig. 
174, commonly designated as steel-plate fans, and those having rotors 




Fig. 174. Standard Steel- 
plate Fan Wheel. 



Fig. 175. " Sirocco " Wheel 
— Turbine Type Impeller. 



Fig. 176. Single Co- 
noidal Fan Wheel. 



with a number of short curved blades, Figs. 175 and 176, generally 
known as multi-vane fans. Both of these types are found in the modern 
power plant though the multi-vane construction is the more common. 
Tall, narrow steel-plate fans are frequently used for induced-draft 



338 



STEAM POWER PLANT ENGINEERING 



work partly because the narrow wheel permits of shorter overhang on 
the fan bearing and partly because they may be operated at low speed 
and are suitable for direct connection to steam engines. Multi-vane 
fans require less space than steel-plate fans of equal capacity and effi- 
ciency and on account of higher speed requirements are more suitable 
for direct connection to electric motors or steam turbines. Each type 
has different characteristics, the nature of which controls the selection 
for a given set of operating conditions. The housings may be ar- 
ranged for top or bottom horizontal discharged, up or down blast, or 
special, depending upon the arrangement of the draft system. 

155. Performance of Fans. — On account of the great number of 
variables involved in the construction and operation of fans simple 
equations or formulas for proportioning the various elements are prac- 
tically impossible. The design of a new fan is largely a matter of trial 
and error based on experiments. For this reason no attempt will be 
made to analyze the problem of design and only such elementary theory 
will be discussed as is necessary for a clear understanding of the prin- 
ciples of operation. 

Pressure. If the delivery pipe of a fan is sealed against discharge 
there is but one pressure in the conduit, namely, static pressure. Re- 
ferring to Fig. 177, A and B represent Pi tot tubes inserted in the dis- 
charge or suction pipe of a centrifugal fan, A being bent to face the 



z;^ 



lamic Opening 



rr^ 



(A) 



c^ 



Static ( B ) 
Opening 



Orifice Closed 



Fig. 177. 



current while B is flush with the inside wall of the casing at right angles 
to it. A receives the full impulse of the stream, and the manometer in- 
dicates the total or dynamic pressure, while B registers the static pressure 
only. With the pipe sealed against discharge, resistance to flow is a 
maximum, there is no flow and the water depression in both manometers 
will be the same, that is, there is only static pressure in the conduit. 

If the discharge orifice is opened to its maximum and there are no 
frictional resistances the static pressure indicated by manometer J5, 



MECHANICAL DRAFT 



339 



Fig. 178, becomes zero while that in A stands at a height equivalent 
to the full impulse of the stream, that is, there is only velocity pressure 
in the conduit. 



^;=^ 



r^ 



/;=^ 



"^ 



a 



(A) 
Orifice Wide Open 



(B) 



Fig. 178. 

If the orifice is partly closed, as in Fig. 179, there will be a water 
depression in both manometers A and 5, that is, there is both velocity 
and static pressure in the conduit. The difference between the de- 
pression in A and B is the pressure due to velocity. By connecting 
the two manometers as indicated in Fig. 179 (C), the velocity pressure 
is given directly. 



/^^ 



T" 

H 

Li 
4# 



(A) 



dy 



(B) 



Orifice partly closed 



r^^ 



ii' 



/7^ 



^^ 



(C) 



(y 



Fig. 179. 

Pressure resulting from the impulse of a current of air flowing at a 
velocity corresponding to that of the tip of the blades is commonly 
designated as the peripheral velocity pressure. 

The ratios between the various pressures are of great importance in 
fan engineering and manufacturers publish characteristic curves showing 
this relationship for various conditions of operation. These charac- 
teristics vary with the type of fan and the design of the blades and hous- 
ing. A few examples are shown in Figs. 180-183. 

The ''ratio of opening," Fig. 180, refers to the actual percentage of 
opening compared v/ith the maximum. The ''ratio of effect" is the 
relative effect produced by restricting the discharge. 



340 



STEAM POWER PLANT ENGINEERING 



Suppose a steel-plate fan with an unrestricted inlet and outlet delivers 
25,000 cu. ft. of air per minute against a dynamic head of 2.14 in. with 
a peripheral velocity requiring 4.5 horsepower. It appears from the 
curves in Fig. 180 that if the discharge outlet is restricted to 50 per cent 
of the full area, only 12,500 cu. ft. will be delivered. The dynamic 
pressure will be increased to 4.28 in., and the power required drops to 
2.7 horsepower. If the outlet be still further reduced to 20 per cent of 
the full opening the capacity will drop to 5000 cu. ft., the pressure 



lOQ 




40 60 80 

Ratio of Effect Pfer cent 

Fig. 180. Characteristic Curves of 22-in. Buffalo Steel-plate Blower (at 

1400 R.P.M.). 

will increase to 4.41 inches, and the power will be decreased to 1.35 
horsepower. With a discharge area of 50 per cent, the mechanical 
efficiency is a maximum, and equal to about 43 per cent. With orifice 
closed the horsepower required to drive the fan is about 24 per cent of 
that required when discharging the maximum volume of air. 

Velocity: In a centrifugal fan operating under constant orifice con- 
ditions and at known air density, the theoretical velocity and pressure 
developed bear a definite relation to the peripheral velocity of the 
fan. For ordinary fan work where air is at a low pressure the 
relationship between pressure and velocity is substantially 



V =V2gh, 



(114) 



MECHANICAL DRAFT 341 

in which 

V = velocity, ft. per sec, 

g = acceleration of gravity; 32.2 (approximately), 
h = head of air causing flow, ft. 

Equation (114) may be reduced to the convenient form 



V = 1096.5 Vp ^ 5, (115) 

in which 

V = velocity, ft. per min., 

p = pressure drop producing velocity, in. of water, 
8 = density of the air, lb. per cu. ft. 

For standard conditions, dry air at 70 deg. fahr. and 29.92 barometer: 

z; = 4005Vp. (116) 

Where quietness of operation is necessary the velocity should be 
limited to 2000 ft. per min. but where this is not essential duct velocities 
as high as 4000 ft. per min. may be used. Since the friction losses of 
a piping system vary with the square of the velocity the usual com- 
promise must be made between size and velocity, otherwise the pressure 
losses become excessive. 

Capacity. For a given fan size, piping system and air density the 
capacity, Q, varies directly as the velocity and hence as the speed of the 
fan, thus, 

Q = vA, (117) 

in which 

Q = volume, cu. ft. per min., 

V = velocity, ft. per min., 
A = area of the conduit. 

Since the velocity varies as the square root of the pressure drop 

Q = KAV^, (118) 

in which 

K = coefficient determined by experiment; other notations as in 
equation (115). 

Horsepower. The horsepower required to operate a fan varies di- 
rectly with the capacity and the total or dynamic pressure, thus: 

_ 5.2 Q X Pa nnnmr;7 Qx-P^ mm 

in which 

E = total efficiency of the blower, 
Pd = dynamic pressure, in. of water. 



342 



STEAM POWER PLANT ENGINEERING 



8 2.8 

2.6 

7 I 2.4 

^2.2 
^6|2.0 

I ='■' 



3rt0 



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BUFFALO TURBO-CONOIDAL 




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NO. 6 

800 R.P.M. 

7080 Peripheral Velocity 






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0.7 ^ 70 

0.6 I 60- 

0.5 -§50 w 

s ^ 
0.4 '^.40 (2 

0.3 Sao w 

0.2 ^20 o 

o e 

0.1 o 10 « 
> 



2 4 6 8 10 12 14 16 

Capacity, Thousands of Cubic Feet per Minute 

Fig. 181. Typical Characteristic Curves of Buffalo "Conoidal" Blower. 



140 

130 

120 

.110 

100 

90 

I 80 

I 70 

60 

50 

40 

30 

20 

10 



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/ 


y 




















\ 




^ 










/ 






















\«1 










/ 
























\\ 








/ 


























\ 









20 



m 



140 



160 



60 80 100 120 

Per cent of Rated Capacity 

Fig. 182. Typical Characteristic Curves of Buffalo "Planoidal" Steel-plate 

Exhausters. 



MECHANICAL DRAFT 



343 



Combining equations (118) and (119) and reducing, remembering 
that for constant orifice conditions and at known air density the veloc- 
ity pressure bears a definite relation to the peripheral velocity, we have 

Hp. = 5pt, (120) 

in which 

B = coefficient involving all constants and reduction factors. 
Equation (120) shows that the horsepower varies as the cube of the 
square root of the pressure. 

150 



140 
130 
120 
110 
100 

§ 90 

u 
o 
^ 80 

70 

60 

50 

40 

30 

20 























..•vf 






-^ 








\ 
















■^A 


v^ 


te^ 










^ 




^ 


\ 


\ 








o.^^ 


^ 


>^^^ 






c..^^ 


^"■^ 










^^ 


-^ 


\ 


^ 


A 


r^ 






,.^-^ 


A 


p 


















/ 


y^ 


^^ 


r::::^ 


:as^ 


J^ 


y 





Perc 


=.nt ot Katea 


Efficiency 




















y 


/ 




^ 


^ 


^ 




=^: 


Per 


'ento 


;Rate 


^ 


^ow;er 








/ 


y 


[^ 


/ 


y 












^eent^ 


£«at 


l£^ 


acitj, 






/ 


/ 


/ 


/ 






i 




















/ 


V 


/ 


/ 


/ 






s 

u 

■2 




















/A 


/ 


/ 


/ 




























/ 




/ 










03 






















/ 


































/ 










— 1 























2 3 4 5 6 

llatio of Static to Velocity Pressure at Fan Outlet. 



Fig. 183. Typical Characteristic Curves of Buffalo "Planoidal" Steel-plate 

Exhausters. 

Since the capacity is directly proportionate to the peripheral velocity 
or fan speed, and the pressure developed varies directly as the square 
of the speed it follows that the horsepower varies as the cube of the 
speed, thus: 

Hp. = M}^\ (121) 

in which 

M = coefficient involving all constant and reduction factors, 
N = speed of the fan, r.p.m. 

The marked increase in power required for even a moderate increase 
in speed should be borne in mind in selecting a fan. It is, as a rule, 
more economical to err in selecting too large a fan than one which 
must be forced above its rated speed. 



344 STEAM POWER PLANT ENGINEERING 

The capacity varies directly with the speed; therefore the horsepower 
also varies with the cube of the capacity. 

Manometric Efficiency. This efficiency is the ratio of the dynamic 
head as actually observed to the maximum theoretical dynamic head, or 

-E^man. = 77 ' (122) 

in which h is determined from the actual manometer reading and H 
is the theoretical maximum head. 

Volumetric Efficiency. This is the ratio between the actual volume 
of air passing in a given time divided by the impeller displacement 
for the same period, or 

-^voi. = ^2jV5» (123) 

in which 
Q = volume discharge, cubic feet per minute, 
D = diameter of the impeller, feet, 
B = width of the impeller, feet, 
N = r.p.m. 

Mechanical efficiency, or simply fan efficiency is the ratio of the 
total work done by the fan in moving the air to the horsepower input 
to the fan, or 

^°^^- ^^1X33,000' ^^^^^ 

in which 

Q = weight discharged, pounds per minute, 
h = dynamic head, feet of air, 
Hi = horsepower input. 

Two efficiencies are sometimes given, (1) that based on the dynamic 
head as in equation (124) and that based on the static head. See 
Fig. 181. 

An analysis of the performance of fans under various operating 
conditions is beyond the scope of this test and the reader is referred 
to the accompanying bibliography for an extended study. 

Measurement of Air in Fan Work: C. H. Treat, Jour. A.S.M.E., Sept., 1912, 
p. 1341. Some Experiences with the Pitot Tube on High and Low Air Velocities: 
F. H. Kneeland, Jour. A.S.M.E., Nov., 1911, p. 1407. Experiments with Ventilating 
Fans and Pipes: Gapt. D. W. Taylor, Soc. Naval Arch, and Marine Engrs., 1905, 
p. 35. The Measurements of Gases: Carl C. Thomas, Jour. Frank. Inst., Nov., 
1911, p. 411. Experiments with the Pitot Tube in Measuring the Velocities of Gases: 
R. Burnham, Eng. News, Dec. 21, 1905, p. 660. Pressure Fans vs. Exhaust Fan: 
Bulletin Am. Inst. Min. Engrs., Feb., 1909. A.S.M.E. 1915 Code for Testing Fans: 
Trans. A.S.M.E., Vol. 37, 1915, p. 1342. 



MECHANICAL DRAFT 345 

156. Selection of Fan. — In general, the multi-vane is more effi- 
cient than the steel-plate fan as ordinarily constructed, and requires 
less space than the latter for equal capacity and efficiency. Another 
important advantage lies in the fact that the higher speed of the multi- 
vane fan permits of direct connection to high-speed prime movers. 
The steel-plate blower, however, is not necessarily a low efficiency 
device since by special design it may be made to give higher efficiencies 
than obtained from the curved short blade construction. Where first 
cost is a consideration and where space Hmitation is of little conse- 
quence the steel-plate fan may be used to advantage. In small plants 
the power requirements for the mechanical draft system are low and 
the type of fan has but little effect on the overall cost of operation, 
but in large central stations the power requirements are considerable 
and the type and attending pressure characteristics greatly influence 
the ultimate economy. 

Having selected the type of fan the first step is to determine the 
size best suited to the required conditions. The influencing factors 
are primarily the volume of air or gas to be delivered and the static 
pressure necessary to overcome the frictional resistances of the system. 
The air requirements and the pressure drop through the boiler equip- 
ment may be calculated as shown in paragraphs 23 and 127. The 
frictional resistance of the air ducts and dampers must be included in 
determining the maximum static pressure. 

The next step is to select from capacity tables (furnished by fan 
builders) the nearest commercial size which will meet the volume and 
static pressure requirements. If the conditions are different from 
those published in the tables the performance under specified condi- 
tions may be approximated by calculation or taken directly from 
''characteristic curves" of the particular type under consideration. 

Two sets of capacity tables are found in practice, the ''rated 
capacity" (such as are reproduced in Tables 67 and 69) and the 
"variable capacity" (one element of which is given in Table 70). 
The former gives the capacity, speed, and horsepower of the different 
fans for various static or total pressures, when operating at what is 
approximately the highest efficiency. These rated capacity tables are 
self-explanatory and require no particular discussion. The variable 
capacity tables give the performance of each size of fan on either side 
of the condition for maximum efficiency and offer practically the same 
information as the characteristic curves. By means of these tables 
the performance of the fan may be readily obtained for practically 
all conditions of operation. 



346 



STEAM POWER PLANT ENGINEERING 



TABLE 67. 

CAPACITIES OF FORCED-DRAFT FANS. 
(Steel Plate Fans.) 



For Forced Draft, Temperature of Air 60°. 















Pressure 


in Inches of Water. 












Cubic Feet 
of Air De- 
































Diam- 
eter of 


livered to 
Furnace 


0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


Fan. 






























per 


S 


Ah 


s 


P^ 


§ 


^ 


S 


aI 


% 


Ah' 


S 


Ah' 


% 


Ah 




Minute. 




W 




W 


Ph 


a 


Ah 


M 


Ah 


W 


Ah 


W 


Ah 


W 


2' 6" 


4,200 


510 


1.6 


560 


1.8 


600 


1.9 


640 


2.1 


710 


2-. 3 


780 


2.5 


850 


2.7 


3' 


5,800 


430 


2.2 


460 


2.4 


490 


2.6 


530 


2.8 


590 


3.1 


640 


3.4 


710 


3.8 


3' G'' 


7,800 


360 


3.0 


400 


3.3 


420 


3.5 


450 


3.8 


500 


4.2 


550 


4.6 


610 


5.1 


4' 


10,000 


320 


3.9 


350 


4.2 


370 


4.4 


400 


4.9 


440 


5.4 


480 


5.9 


530 


6.5 


4' 6' 


12,400 


290 


4.8 


310 


5.2 


330 


5.6 


360 


6.0 


400 


6.7 


430 


7.3 


470 


8.0 


5' 


15,200 


250 


5.9 


270 


6.4 


290 


6.8 


310 


7.4 


350 


8.2 


380 


8.9 


420 


9.8 


5' ^'' 


18,200 


230 


7.0 


250 


7.7 


270 


8.2 


300 


8.8 


330 


9.8 


360 


10.6 


390 


11.8 


6' 


21,400 


210 


8.3 


230 


9.1 


250 


9.6 


260 


10.4 


290 


11.5 


320 


12.5 


350 


13.9 


r 


28,800 


180 


11.2 


200 


12.2 


210 


13.0 


230 


14.0 


250 


15.5 


280 


16.8 


300 


18.7 


8' 


37,200 


160 


14.4 


170 


15.7 


190 


16.7 


200 


18.1 


220 


20.1 


240 


21.8 


270 


22.5 


9' 


46,800 


140 


18.1 


160 


19.8 


170 


21.1 


180 


22.7 


200 


25.3 


220 


27.4 


240 


30.3 


10' 


57,400 


130 


22.2 


140 


24.3 


150 


25.8 


160 


27.9 


180 


3.1 


200 


33.6 


210 


37.2 



Discharge velocity 2000 feet per minute. 

TABLE 68. 

CAPACITIES OF INDUCED-DRAFT FANS. 

(Steel Plate Fans.) 









For Induced Draft, Temp. 


of Flue Gases 500°. 












Cubic Feet 






Pressure 


in Inches of Water. 




of Air at 












Diam- 


60°Temp. 


0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


eter of 


Drawn 
into Fur- 
nace per 
















Fan. 


H 


Ah* 


s 


Ah' 


S' 


Ah* 


N 


Ah* 


S 


P^ 


S 


Ah* 


S 


Ah' 




Minute. 


Ah 

688 


w 

2.2 


Ph 

756 


2.4 


Ah 

810 


w 

2.6 


Ah 

864 


2.8 


958 


M 
3.1 


Ah 

1053 


W 

3.4 


Ah 

1147 


m 


2' 6" 


3,000 


3.6 


V 


4,200 


580 


3.0 


621 


3.2 


661 


3.5 


715 


3.8 


796 


4.2 


864 


4.6 


958 


5.1 


V 6^ 


5,700 


486 


4.0 


540 


4.5 


567 


4.7 


607 


5.1 


675 


5.7 


742 


6.2 


823 


6.9 


4' 


7,300 


432 


5.3 


472 


5.7 


500 


6.1 


540 


6.6 


594 


7.3 


648 


8.0 


715 


8.8 


4' 6" 


9,300 


390 


6.5 


418 


7.0 


445 


7.5 


486 


8.1 


540 


9.0 


580 


9.8 


634 


10.8 


5' 


11,100 


337 


8.0 


364 


8.6 


391 


9.2 


418 


10.0 


472 


11.1 


513 


12.0 


567 


13.2 


5' 6" 


13,300 


310 


9.5 


337 


10.4 


364 


11.1 


405 


11.9 


445 


13.2 


486 


14.3 


526 


15.9 


6' 


1S600 


283 


11.2 


310 


12.3 


337 


13.0 


351 


14.0 


391 


15.5 


432 


16.9 


472 


18.7 


r 


.000 


243 


15.1 


270 


16.5 


283 


17.5 


310 


18.9 


337 


20.9 


378 


22.6 


405 


25.2 


8' 


27,100 


216 


19.4 


230 


21.2 


256 


22.: 


270 


24.4 


297 


27.1 


324 


29.4 


364 


30.4 


9' 


34,200 


189 


24.4 


216 


26.7 


230 


28.5 


243 


30.6 


270 


34.1 


297 


37.0 


324 


40.9 


10' 


41,900 


175 


30.0 


190 


32.8 


202 


34.8 


216 


37.6 


243 


41.8 


270 


45.3 


283 


50.2 



MECHANICAL DRAFT 



347 



TABLE 69. 

CAPACITIES OF FORCED-DRAFT FANS.* 
(Sirocco Type.) 

(Figures given Represent Dynamic Pressures in Ozs. per Sq. In. For Static Pressure Deduct 28.8 Per Cent. 
For Velocity Pressure Deduct 71.2 Per Cent.) 



.2^ 




iOz. 


h Oz. 


JOz. 


lOz. 


UOz. 


UOz. 


IJ Oz. 


2 0z. 


2^0z. 


3 0z. 


Si 
























6 


Cu. Ft. 
R.P.M. 
B.H.P. 


155 

1,145 

0.0185 


220 
1,615 
0.052 


270 
1,980 
0.095 


310 
2,290 
0.147 


350 
2,560 
0.205 


380 
2.800 
0.270 


410 
3,025 
0.34 


440 
3,230 
0.42 


490 
3.616 
0.58 


540 
3,960 
0.76 


12 


Cu. Ft. 
R.P.M. 
B.H.P. 


625 

572 

0.074 


880 

808 

0.208 


1,080 

990 

0.381 


1,250 
1,145 

0.588 


1,400 
1,280 
0.82 


1,530 
1,400 
1.08 


1.650 
1.512 
1.36 


1.770 
1,615 
1.66 


1.970 
1,808 
2.32 


2.170 
1.980 
3.05 


18 


Cu. Ft. 
R.P.M. 
B.H.P. 


1,410 

381 

0.167 


1,990 

538 

0.470 


2,440 

660 

0.862 


2,820 
762 
1.33 


3,160 
850 
1.85 


3,450 

933 

2.43 


3,720 
1.010 
3.07 


3.980 
1,076 
3.75 


4,450 
1,204 
5.25 


4.880 

1,320 

6.9 


24 


Cu. Ft. 
R.P.M. 
B.H.P. 


2,500 

286 

0.296 


* 3,540 

404 
0.832 


4,340 
495 
1.53 


5,000 
572 
2.35 


5,600 

640 

3.28 


6,120 

700 
4.32 


6,620 

756 

5.44 


7,080 

807 

6.64 


7,900 
904 
9.3 


8.680 
990 
12.2 


30 


Cu. Ft. 
R.P.M. 
B.H.P. 


3,910 

228 
0.460 


5.520 
322 
1.30 


6,770 
395 
2.40 


7,820 

456 

3.68 


8,750 

510 

5.15 


9,600 

560 

6.75 


10,350 

604 

8.53 


11,050 
645 
10.4 


12,350 
722 
14.5 


13,550 
790 
19.1 


36 


Cu. Ft. 
R.P.M. 
B.H.P. 


5,650 

190 

0.665 


7,950 
269 
1.87 


9,750 

330 

3.44 


11,300 
381 
5.30 


12,640 
425 

7.40 


13,800 

466 

9.72 


14,900 

504 

12.25 


15,900 
538 
15.0 


17,800 

602 

20.9 


19,500 
660 
27.5 


48 


Cu. Ft. 
R.P.M. 
B.H.P. 


10,000 

143 

1.18 


14,150 

202 

3.32 


17,350 

248 

6.10 


20,000 

286 

9.40 


22,400 
320 
13.1 


24,500 
350 
17.2 


26,500 

378 

21.75 


28,300 

403 

26.6 


31.600 
452 
37.1 


34,700 

495 

48.8 


60 


Cu. Ft. 
R.P.M. 
B.H.P. 


15,650 

114 

1.84 


22,100 

161 

5.20 


27,100 

198 

9.58 


31,300 
228 
14.7 


35,000 

255 

20.6 


38,400 

280 

27.0 


41,400 

302 

34.1 


44.200 

322 

41.6 


49.400 
361 
58.2 


54.200 
396 
76.5 


72 


Cu. Ft. 
R.P.M. 
B.H.P. 


22,600 

95 

2.66 


31,800 
134 

7.48 


39,000 

165 

13.7 


45,200 

190 

21.2 


50,600 
212 
29.6 


55,200 

233 

38.9 


59,600 

252 

49.0 


63.600 

269 

59.8 


71.200 

301 

83.6 


78.000 
330 
110 


84 


Cu. Ft. 
R.P.M. 
B.H.P. 


30,800 

81 

3.61 


43,400 

115 

10.2 


53,200 

142 

18.7 


61,600 

163 

28.9 


68,700 

182 

40.4 


75,200 

200 

53.0 


81,200 

216 

66.8 


86,800 

231 

81.7 


97,100 
258 
114 


106,400 
283 
150 


90 


Cu. Ft. 
R.P.M. 
B.H.P. 


35,250 

76 

4.14 


49,800 

107 

11.7 


61,000 

132 

21.5 


70,500 

152 

33.1 


78,800 

170 

46.2 


86,400 

186 

60.7 


93.300 
201 
76.7 


99.600 

214 

93.6 


111,200 
241 
131 


122.000 
264 
172 



A number of sizes have been omitted. 



157. Chimney vs. Mechanical Draft. — The choice of chimney or 
mechanical draft depends largely upon local conditions. Where there 
are no hmitations to the height of stack mechanical draft offers many 
advantages over chimney draft. With certain types of grates and 
for low-grade fuels and anthracite culm or dust, it is indispensable. 
Again, where a fair quahty of fuel is ol)tainable the size of plant may 
determine the choice. 

First Cost: In small plants of, say, 100 to 150 horsepower .he cost 
of a guyed steel chimney, 75 feet in height or less, would cost practi- 
cally nothing for operation, while the power required to operate a fan 



348 



STEAM POWER PLANT ENGINEERING 



in so small a plant would amount to 5 per cent or more of the total 
steaming capacity. 

A tall, self-supporting chimney for larger plants, however, is very 
costly as compared with a fan system of equal capacity. For example, 
a brick chimney 175 feet high and 10 feet in diameter, foundation and 
all, capable of furnishing the necessary draft for a 3000-horsepower 
plant, will cost about $10,000. A two-fan induced system of equiva- 
lent capacity will cost in the neighborhood of $5000, a one-fan system 
$3500, and a forced-draft system $2500. See Fig. 179. With interest 
at 5 per cent, depreciation 5 per cent, taxes 1 per cent, and insurance 
one-half per cent, the annual fixed charges will be $575, $402.50, $287.50 
respectively, for the fan equipment. 













TA] 


BLE 


70. 


















TYPICAL VARIABLE CAPACITY CHART. 














Performance of No. 6 Buffalo Turbo-Conoidal Fan. 








Outlet 
Veloc- 


Capac- 
ity, 

Cu. Ft. 
Per 
Min. 


Add 
for 
Total 
Pres- 
sure. 


Static Pressure, Inches of Water. 


ity, 
Ft. Per 


* 


1 


U 


' 


3 


Min. 


R.p.m. 


Hp. 


R.p.m. 


Hp. 


R.p.m. 


Hp. 


R.p.m. 


Hp. 


R.p.m. 


Hp. 


1000 


5,250 


0.063 


415 


0.71 


















1200 


6,300 
7,350 


0.090 
0.122 


443 
472 


0.94 
1.23 


563 

587 


1.60 
1.99 


663 

682 


2,29 
2.76 


748 
767 


3,13 
3.56 






1400 


910 


5.57 


1600 


8,400 


0.160 


503 


1.56 


613 


2.43 


705 


3.31 


785 


4.19 


930 


6.09 


1800 


9.450 


0.202 


535 


1.96 


642 


2.93 


730 


3.92 


808 


4.93 


947 


6.88 


2000 


10,500 


0.250 


562 


2.41 


670 


3.51 


757 


4.61 


833 


5.71 


967 


7.91 


2200 


11,550 


0.302 


602 


2.94 


702 


4.19 


785 


5.37 


858 


6.58 


988 


9.01 


2400 


12,600 


0.360 


637 


3.53 


733 


4.93 


813 


6.23 


885 


7.54 


1013 


10.19 


2600 


13,650 


0.422 


670 


4.21 


765 


5.76 


843 


7.18 


915 


8.61 


1040 


11.45 


2800 


14,700 


0.489 


708 


5.01 


798 


6.69 


873 


8.24 


945 


9.76 


1067 


12.85 


3000 


15,750 


0.560 


745 


5.87 


832 


7.71 


907 


9.42 


973 


11.09 


1093 


14.29 


3200 


16,790 


0.638 










970 


12.06 


1037 


14.02 


1152 


17.71 















This chart has been considerably condensed. In the original table static pressures and velocities are 
given for a wider range and at narrower increment. 



Depreciation and Maintenance: The depreciation of a well-designed 
masonry or concrete stack is very low, and 2 per cent is a hberal factor. 
Maintenance is practically negligible, as it requires no attention what- 
ever for years. A steel stack, however, must be kept well painted or 
corrosion will take place rapidly. The depreciation and maintenance 
charges on a mechanical-draft system will range from 4 per cent to 10 
per cent of the original outlay. 

Cost of Operation: Once erected, the comparative cost of operating 



MECHANICAL DRAFT 



349 



a chimney is practically nothing; that is, of course, on the assumption 
that the chimney and fan exhaust equal volumes of gas per pound of 
fuel and at the same temperature. A fan system requires for its oper- 
ation from one and one-half per cent to five per cent of the total steam- 
ing capacity of the plant, depending upon the type and character of 
the fan engine or motor, and the conditions of operation. 



15 




























































































































14 






























































13 


























































































































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1000 



2000 



3000 



Horse Power 
Fig. 184. Comparative Costs of Chimneys and Mechanical Draft (W. B. Snow). 

Efficiency: With fan draft a very thick fire can be maintained on 
the grate, thus permitting a high rate of combustion, and minimum 
air per pound of fuel, both of which result in increased boiler efficiency. 
The influence of the rate of combustion on air supply in a specific case 
is illustrated in Fig. 185. For the same temperature of discharge each 
pound of air in excess of theoretical requirements results in a loss of 
about one per cent of the total heat in the fuel. With fan draft an 
average figure is 18 pounds of air per pound of bituminous coal against 
24 pounds for the chimney, a saving of 5 per cent in favor of the fan. 



350 



STEAM POWER PLANT ENGINEERING 



Again, a fan permits of a low temperature of the flue gases without 
affecting the draft, while lowering the temperature in the chimney 
reduces the draft as shown in Table 36. From Table 14 we see that 
a reduction in flue gas temperature of 25 deg. fahr. will increase the 



^200 



100 



o 

O 



aO 30 

Lb.Coal Burned Per Sq.Ft.Grate Per Hr. 



50 



Fig. 185. Influence of Rate of Combustion on Air Supply — • Forced Draft. 



boiler efficiency about one per cent. With an economizer the flue 
gases may be reduced to 350 deg. fahr., with a net saving of about 
500 — 350 = 150, or 6 per cent of the total fuel. It is in this con- 
nection that the fan draft is peculiarly suitable. Of course, the chim- 
ney may be provided with an economizer, effecting the same reduction 
in temperature, but its height must be made sufficiently great to over- 
come the additional resistance of the economizer and the reduction in 
temperature of the chimney gases. 

Flexibility : With a fan the draft may be readily regulated for sudden 
increased or decreased requirements, independent of the boiler per- 
formance. Damp and muggy days appreciably affect the draft of a 
chimney, as do adverse air currents and high winds. 

Smoke: Smokeless combustion is more readily effected with arti- 
ficial draft than with natural draft, as a thicker fire can be carried, and 
the correct proportion of air can be more readily adjusted. 

Advantages of Forced Draft on Peak Loads: Elec. Wld., July 8, 1916, p. 68; Sept. 
16, 1916, p. 583. 
Notes on Fans: Power, June 15, 1915, p. 816. 



MECHANICAL DRAFT 351 



PROBLEMS. 

L Dry air is flowing through a conduit, the velocity head (as indicated in Fig. 
179) being one inch water. If one cu. ft. air weighs 0.074 lb., required the velocity 
in ft. per sec. 

2. Let the cross-sectional area of the conduit in Prob. 1 be 2 sq. ft. and the static 
pressure 0.5 in water. Required the output horsepower of the fan. 

3. It is required to supply 20,000 cu. ft. air per min. to a furnace under a static 
pressure of one inch water. The cross-sectional area of the conduit is 3.33 sq. ft. 
The mechanical efficiency of the fan is 40 per cent. One cu. ft. of air weighs 0.074 
lb. Calculate the horsepower required to drive the fan. 

4. Required the horsepower necessary to operate the fan in Problem 3 if the static 
pressure is increased in 2 in. water. 

5. If the rated speed of fan in Problem 3 is 2000 r.p.m. required the horsepower if 
the speed is increased to 4000 r.p.m. 

6. The demands on a fan running 2000 r.p.m. have increased and it is estimated 
the fan will deliver the required volume of air if speed is increased to 4000 r.p.m. 
Show why it will be much more economical to replace blower with one designed to 
deliver the required volume. 

7. Required the capacity of an induced fan suitable for the conditions stated 
in Problem 3, Chapter VI. 

8. Required the power necessary to operate the fan in Problem 7, if its mechanical 
efficiency is 60 per cent. 



CHAPTER IX 

RECIPROCATING STEAM ENGINES 

158. Introductory. — The type of prime mover best suited for a 
given installation is the one which delivers the required power at the 
lowest cost, taking into consideration all charges, fixed and operating. 
These include not only the cost of fuel, labor, supplies and repairs, 
but all overhead charges such as interest on the investment, deprecia- 
tion, maintenance and taxes. Space requirements and continuity of 
operation are often of vital importance, and may greatly influence 
the selection of type of prime mover and auxiliary apparatus. In 
many situations the gas engine and producer are productive of the 
highest commercial economy; in others the choice lies between the 
reciprocating steam engine or turbine, occasionally the hydroelectric 
plant offers the best returns, but each proposed installation is a problem 
in itself, and general rules are without purpose. 

The reciprocating steam engine is the most widely distributed prime 
mover in the power world, and although its field of usefulness has been 
greatly encroached upon in recent years by the steam turbine and gas 
engine it is still an important heat engine and will probably continue 
to be a factor for years to come. In a general sense the piston engine 
is superior to the turbine for variable speed, slow rotative speeds and 
heavy starting torque, while the turbine has superseded the engine 
for large central station units and for auxiliaries requiring high ro- 
tative speed. The high-speed turbine in connection with efficient 
reduction gearing has many advantages over the piston engine for 
low-speed drives and is rapidly replacing the latter in this connection. 
From a purely thermal standpoint the Diesel type of internal com- 
bustion engine is superior to the steam engine and the turbine is more 
economical in space requirements, but taking into consideration all 
of the items affecting the production of power, the reciprocating engine 
may still prove to be the better investment in many situations, at 
least for sizes under 1000 horsepower. 

Improvement in the heat efficiency of the piston engine within the 
past three years has been remarkable and single cylinder units are 
being operated with steam consumptions lower than that obtained 
from the older types of compound units. A few years ago the piston 
engine appeared to be doomed to the scrap heap but the unusual econ- 

352 



RECIPROCATING STEAM ENGINES 353 

omies effected in the later designs has made it once more a torniidal;le 
competitor of the steam turbine, at least for moderate power require- 
ments and non-condensing service. 

Present Status of Prime Movers: Pro. A.I.E.E., June, 1914, p. 953; Jan., 1915, p. 102. 

159. The Ideal Engine. — In every heat engine the working fluid goes 
through a circuit or cycle of operation. Beginning at a particular 
condition it passes through a series of successive states of pressure, 
volume and temperature and returns to the initial condition. An 
ideally perfect engine which effects the highest possible conversion 
of heat into mechanical work for a given cycle is taken as a standard 
of comparison for the performance of the actual engine. There are 
several cycles which simulate more or less the action of steam in the 
actual engine, but the Rankine cycle meets the conditions of most 
engines and for that reason has been adopted as a standard. The 
various cycles are treated at length in Chapter XXIV and need not 
be considered here. 

In order to realize the ideal Rankine cycle the walls of the cylinder 
and the piston must be non-conducting, expansion after cut-off must 
be adiabatic and carried down to the existing back pressure, the action 
of the valves must be instantaneous and steam passages must be 
sufficiently large to prevent wire drawing. None of these conditions 
is fulfilled by the actual engine. The various losses which prevent 
the actual engine from obtaining the efficiency of the ideal are outlined 
in paragraphs 169 to 177. 

The heat supplied, heat consumption, efficiency and water rate of 
a perfect engine operating in the Rankine cycle are treated at length 
in Chapter XXIV and may be summed up as follows: 

Heat supplied = Hi — qn, B.t.u., (126) 

Heat absorbed = Hi — Hn, B.t.u., (127) 

Efficiency, Er = tt -' (128) 



Hi - On 

2546 
Hi- H, 



2546 
Water rate, Wr = 77 tj- , lb. per hp-hr., (129) 



in which 

Hi = initial heat content of the steam, 

Hn = final heat content after adiabatic expansion from initial con- 
dition to final condition n, 
Qn = heat of the hquid corresponding to exhaust temperature. 

The average engine seldom expands to the existing back pressure 
and though this is a fault chargeable against it, occasion may arise to 



354 



STEAM POWER PIANT ENGINEERING 



ma 


^ 






p *" 


1 rrw=^' 


u 




: ,.- 


-J 


P-) 


1J 


/' ->^ 


_^-^^^_.^ 


r 


^ Ti 


// 


1^;^ 






M — ' 


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O 

bC 



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'fab 
g 



1=1 



RECIPROCATING STEAM ENGINES 



355 



compare the actual cycle with the theoretical in which expansion is 

not complete. The various theoretical quantities for this condition of 

incomplete expansion (see paragraph 462) may be calculated as follows: 

Heat supplied = Hi — Qn B.t.u. per lb. (130) 

It will be noted that this is the same as for complete expansion. 

Heat absorbed = Hi- Hc + jh {Pc - P2) Vo B.t.u. per lb., (131) 

Hi - He-\- j}^ {Pc - P2) Ve 



Efficiency, Er' 

Water rate, W/ 
in which 



Hi - Qn 

2546 



Hi-He + j\^{Pc-P2)Vc 



lb. per hp-hr. 



(132) 
(133) 



He = heat content at release pressure Pc after adiabatic expansion, 
Pc = release pressure, lb. per sq. ft., 

Vc = specific volume of the fluid under release conditions, cu. ft. 
per lb. (Other notations as for complete expansion.) 



TABLE 71. 

THEORETICAL EFFICIENCIES AND WATER RATES OF PERFECT ENGINES 
OPERATING IN THE CARNOT AND RANKINE CYCLES. 

(Saturated Steam.) 





Condensing.* 


Non-condensing. 


Initial 
















Pressure, 


Efficiency. 


Water Rate. 


Efficiency. 


Water Rate. 




















c 


R 


C 


R 


C 


R 


C 


R 


50 


27.18 


24.98 


10.13 


8.98 


9.32 


8.98 


29.56 


28.51 


100 


31.51 


28.47 


9.10 


7.85 


14.70 


13.88 


19.48 


18.22 


150 


34.10 


30.60 


8.65 


7.26 


17.90 


16.65 


16.46 


15.08 


200 


35.91 


31.88 


8.41 


6.94 


20.19 


18.60 


14.94 


13.44 


250 


37.34 


32.93 


8.25 


6.70 


21.97 


20.05 


14.02 


12.42 


300 


38.51 


33.76 


8.14 


6.52 


23.42 


21.22 


13.39 


11.71 


400 


40.37 


35.10 


8.00 


6.25 


25.74 


23.07 


12.53 


10.73 


500 


41.79 


36.06 


7.988 


6.07 


27.54 


24.46 


12.12 


10.10 


600 


43.00 


36.84 


7.987 


5.94 


29.02 


25.57 


11.83 


9.66 



* Absolute back pressure 0.5 lb. per sq. in. 

Direct-acting steam pumps and engines taking steam full stroke 
have the following theoretical possibilities (see paragraph 463) : 

Heat supphed = Hi — Qn B.t.u. per lb., (134) 

Heat absorbed = yj^ (Pi — P2) Vi B.t.u. per lb., (135) 

(Pi - P2) V, 



Efficiency, E/' = 
Water rate, Wr" = 



778 {Hi - qn) 
2546 X 778 
{Pi - P2) V, 



(136) 
(137) 



356 



STEAM POWER PLANT ENGINEERING 



in which 

Vi = specific volume of the steam at pressure Pi, cu. ft. per lb. 

(Other notations as above.) 

160. Efficiency Standards. — The performance of the actual engine is 
variously stated as follows: 

1. Steam consumption, pounds per hour or hp-hr. 

2. Heat consumption, B.t.u. per hp-hr. or per hp. per minute. 

3. Thermal efficiency, per cent. 

4. Mechanical efficiency, per cent. 

5. Rankine cycle ratio, per cent. 

6. Cylinder efficiency, per cent. 

7. Commercial efficiency. 

8. Duty. 




Fig. 187. Typical Piston Engine, Single Cylinder, Automatic Governor. 



RECIPROCATING STEAM ENGINES 357 

The indicator offers the simplest means of measuring the output of 
a piston engine, and for this reason the performance is usually stated in 
terms of indicated horsepower. The indicated horsepower is always 
greater than the net available power by an amount equivalent to the 
friction of the mechanism. The power actually developed, or brake 
horsepower, is not readily obtained except for small sizes, and it is 
customary to approximate this value by deducting the indicated horse- 
power when running idle from the indicated horsepower when running 
under the given load. This does not give the true effective power, 
but is suifficiently accurate for most commercial purposes. (See para- 
graph 174.) The output of steam turbines and piston engines driving 
electrical machinery is conveniently stated in electrical horsepower or 
kilowatts, since the electrical measurements are readily made. The 
electrical output as measured at the switchboard gives the net effective 
work, and automatically deducts the machine losses. Large turbines 
are usually tested at the factory by means of suitable water brakes, 
and the brake horsepower may be obtained from the makers. 

161. Steam Consumption or Water Rate. — The most generally used 
measure of the performance of a steam engine is the steam consump- 
tion per hour or per unit of work output. Since the indicator offers 
the simplest means of measuring the output the performance is usually 
stated in terms of indicated horsepower."^ By plotting the total weight 
of steam passing through the engine as ordinates and the indicated 
horsepower as abscissas the performance of the engine at all loads 
may be seen at a glance. Just what form the total water rate curve 
will take depends largely on the type of governing and the form of 
valve gear. If the control is by throttling, the total water rate curve 
is substantially a straight line, and the relation is commonly called 
the Willans line, Fig. 214. When two points on this hne, or one point 
and the slope, are given the hne can be drawn at once. If the control 
is by ''cutting off" the curve departs somewhat from a straight line, 
but in many cases the departure is insignificant. Fig. 222. Dividing 
the ordinates by corresponding abscissas gives the steam consumption 
per indicated horsepower-hour or unit water rate. Fig-. 214. 

For electrically driven machinery the economy is given as steam 
consumption per electrical horsepower-hour or per kilowatt-hour. A 
study of the unit water rate curve will show that the steam consump- 
tion decreases with the load up to a certain point and then increases. 

* This must not be confused with the steam accounted for by the indicator diagram, 
or, as it is commonly called, the indicated steam consumption. The former refers to 
the actual weight of fluid flowing through the cyhnder and the latter to the weight 
of steam calculated from the indicator card. (See paragraph 8, Appendix B.) 



» 



358 STEAM POWER PLANT ENGINEERING 

This point of minimum steam consumption corresponds ordinarily to 
the rated load. If the initial pressure, quality and back pressure 
were constant for all conditions of operation the water rate would be 
a true measure of heat efficiency, but since this is not the case the 
water rate under actual conditions is of little value in comparing per- 
formances. The water rate may be used as a means of comparison 
provided suitable corrections are made for pressure and quality, but 
this procedure is not common. The water rates for different types 
of engines are given further on in the chapter. 

162. Heat Consumption. — Because of the extreme variation in 
steam conditions the performance of all engines and turbines is best 
expressed in terms of the heat consumption per unit output measured 
above the maximum theoretical temperature at which the condensation 
can be returned to the boiler. This temperature is called the ideal 
feed-water temperature. Thus the ideal feed-water temperature of an 
unjacketed non-condensing engine without receiver coils exhausting at 
standard atmospheric pressure is 212 deg. fahr., and that of a con- 
densing engine exhausting against an absolute back pressure of two 
pounds is 126 deg. fahr. If the engine is fitted with jackets and re- 
heating coils the heat of the hquid at jacket and coil pressure should 
be added to that of the exhaust in determining the ideal feed-water 
temperature. For example, if a condensing engine exhausts against an 
absolute back pressure of two pounds, and ten per cent of the total 
weight exhausted is condensed in the jackets under a pressure of 150 
pounds absolute, the ideal feed- water temperature will be 159.5 deg. 
fahr. (Heat of the liquid at 150 pounds absolute = 330 B.t.u. per 
pound. Heat added by the jackets to the feed water = 330 X 0.1 = 33. 
Heat of the liquid at two pounds absolute = 94 B.t.u. 94 + 33 = 127 
B.t.u., which corresponds to an actual temperature of 159.5 deg. fahr.) 

Example 25. (1) A compound condensing engine develops one brake 
horsepower-hour on a steam consumption of 8.5 pounds, initial pres- 
sure 200 pounds absolute, superheat 250 deg. fahr., exhaust pressure 
0.5 pound absolute, release pressure two pounds absolute. (2) The 
same engine when using wet steam develops one brake horsepower- 
hour on a steam consumption of 12 pounds per hour, initial pressure 
150 pounds absolute, quality 98 per cent, exhaust pressure two pounds 
absolute, release pressure four pounds absolute. 

Determine the comparative heat consumption of the two engines. 

Superheated steam engine : 

Hi = 1332approx. (from steam tables), 

Qn = 48, 

Heat supplied per br. hp-hr. = 8.5 (1332 - 48) = 10,914 B.t.u., 
Heat supplied per br. hp. per minute = 181.9 B.t.u., 



RECIPROCATING STEAM ENGINES 



359 




«fl H 



360 STEAM POWER PLANT ENGINEERING 

Saturated steam engine: 

Hi = xin + qi 

= 0.98 X 863.2 + 330.2 = 1176.1, 

(This may be obtained directly from the Mollier diagram.) 
qn = 94, 

Heat suppUed per br.hp-hr. = 12 (1176.1 - 94) = 12,985 B.t.u., 
Heat supplied per br.hp. per minute = 216.4. 
Economy of superheated steam 

-|2 § 5 

(1) in steam consumption, 100 = 29.2 per cent, 

(2) in heat consumption, 100 ' — '— = 15.9 per cent. 

The heat consumption for different types of piston engines is given 
further on in the chapter. 

163. Thermal Efficiency. — The thermal efficiency of a steam engine 
or turbine is the ratio of the heat converted into useful work to that 
supplied, measured above the heat of the liquid at exhaust steam 
temperature.* If the heat consumption is expressed in terms of i.hp-hr., 
the ratio becomes the indicated thermal efficiency. Since the heat 
equivalent of one horsepower, using the latest accepted values, is 
42.44 B.t.u. per minute or 2546 B.t.u. per hour, this relationship may 

be expressed 

2546 



B.t.u. suppneu per or. up-ur. 

(138) 



in which 



W {Hi - qn) 



W = the weight of steam supplied, pounds per developed horse- 
power-hour. 
Hi and qn as in equation (129). 

If measured in electrical units this relationship becomes 

in which 

Wi = pounds per kilowatt-hour; other notations as in (129). 

* The heat supplied is often measured above the actual feed-water temperature 
but the latter is not dependent upon the performance of the engine and hence is not 
satisfactory for purposes of comparison. 



RECIPROCATING iSTEAM ENGINES 



361 




■^ 



! C! 

i w 



o 

a 

o 

O 



±^ a 

d 



I 



362 



STEAM POWER PLANT ENGINEERING 



Example 26. Determine the thermal efficiencies for the two engines 
using the data of the preceding example. 
Superheated steam engine, 

2546 
^' = 8.5(1332-48) = ""^^^ = 23.3 per cent. 

Saturated steam engine, 
2546 



E,= 



12(1176.1 - 94) 



0.196 = 19.6 per cent. 



The thermal efficiency of the actual engine varies from 5 per cent 
for the poorer grades of non-condensing engines to 25 per cent for the 
best recorded performance to date. As far as thermal efficiency is 
concerned the piston engine still leads the turbine for sizes under 
2000 horsepower. 

164. Mechanical Efficiency. — The ratio of the developed or brake 
horsepower to the indicated power is the mechanical efficiency of the 
engine; the ratio of the electrical horsepower to the indicated power 
is the mechanical efficiency of the engine and generator combined; 
and the ratio of the pump horsepower to the indicated power of the 
engine is the mechanical efficiency of the engine and pump combined. 
The percentage of work lost in friction is therefore the difference be- 
tween 100 per cent and the mechanical efficiency in per cent. (See 
also paragraph 174.) 

TABLE 72. 

MECHANICAL EFFICIENCIES OF ENGINES. 



Kind of Engine. 


Horse Power. 


Efficiency at Full 
Load. 


Simple : 

1. High-speed, non-condensing 


150 
170 
275 

150 

160 

900 

1000 

5500 

7500 

865 
712 


95 5 


2. High-speed, condensing 


96 


3. Low-speed, non-condensing 


94 


Compound: 

4. High-speed, non-condensing 


94 


5. High-speed, condensing 


98 


6. Low-speed, non-condensing 


95 


7. Low-speed, condensing 


95 


8. Do 


95 2* 


9. Do 


93.0* 


Triple: (combined efficiency of engine and 
pump) 
10. Pumping engine 


97 4 


Quadruple: (combined efficiency of engine and 
pump) 
IL Pumping engine 


93 







• Combined efficiency of engine and generator. 



RECIPROCATING STEAM ENGINES 



363 



The mechanical efficiency of piston engines at rated load varies from 
85 per cent for the cheaper grades of engines to 95 per cent and even 
98 per cent for the better types. The size of engine has practically 
no influence on the mechanical efficiency, though the smaller machines 
are apt to have a lower efficiency because of the cheaper construction. 
Generator efficiencies at full load vary from 86 per cent for the 15- 
kilowatt size to 94 per cent for units of 2000 kilowatts rated capacity. 
The generator efficiency of very large turbo-alternators, 25,000 kilo- 
watt rated capacity or more, is in the neighborhood of 96 per cent. 
The overall or combined efficiency at rated load varies from 75 per 
cent for small units to 93 per cent for larger ones. A few examples 
of high engine efficiency are cited in Table 72. The efficiency at 
fractional loads for a specific case are illustrated in Fig. 190. 



























n 




£ 








-f 


^ ^ 






^ 




















(^ 


^f] 




^ 


'^O 


o 




















/ 


f 


y 


.^ 


^ 


^ 




















/ 


/ 


.^ 






















/ 


/ 




v1 








Mechanical 

EflBciencies of 

75 K.W.Geuerating Set 

Engine, Simple High Speed 

Ison Condensing 






A 


V 

p 


/ 














r 




r 














/ 


/ 































/ 



























10 20 



50 60 70 80 90 100 
Per Cent of Rated Load 

Fig. 190. 



110 1^ 130 140 150 



Lucke, Engineering Thermodynamics, p. 370, states that the mechan- 
ical efficiency of the piston engine is independent of the speed and 
that it may be expressed 

Em = l- K,- -^, (140) 

m.e.p. 

in which 

Ki = constant, varying from 0.02 to 0.05, 
K2 = constant, varying from 1.3 to 2.0, 
m.e.p. = mean effective pressure, lb. per sq. in. 



165. Rankine Cycle Ratio. — The degree of perfection of an engine 
or the extent to which the theoretical possibilities are realized is the 
ratio of the thermal efficiency of the actual engine to that of an ideally 
perfect engine working in the Rankine cycle with complete expansion. 



364 STEAM POWER PLANT ENGINEERING 

This is called the Rankine cycle ratio or potential efficiency* It is 
the accepted standard for comparing the performance of steam engines 
and steam turbines. 

If E = Rankine cycle ratio t 

Et = thermal efficiency of the actual engine, 

Er = efficiency of the ideal engine work in the Rankine cycle with 
complete expansion. 

Then E = ^' (141) 

From equation (138) 

And from equation (128) 

Whence 



Et = ^^^^ 



Er 



W (Hi - qn) 

Hi — Hn 



W {Hi - Qn) Hi- Qn 

'''' (143) 



W {Hi - Hn) 

Example 27. Determine the Rankine cycle ratio of the two engines 
specified in paragraph 162. 
Superheated steam engine : 

^ = 8.5 (1332 - 908) = 0-706 = 70.6 per cent. 

Saturated steam engine: 

9^4fi 
^ = l2(rm^898) = 0-^*^3 = 76.3 per cent. 

Tables 81 and 83 give the best recorded Rankine cycle ratios for 
current practice. 

166. Cylinder Efficiency. — The piston engine seldom expands the 
steam down to the existing back pressure but releases from two to five 
pounds above this point in condensing engines and from 15 to 20 pounds 
above in non-condensing engines. The ideal cycle corresponding to 
this condition is the Rankine cycle with incomplete expansion. The 
ratio of the thermal efficiency of the actual engine to that of the ideal 
engine working in the incomplete cycle is a true measure of the degree 

* The term "thermodynamic efficiency" or "efficiency" without qualification is 
ordinarily interpreted as the Rankine cycle ratio though some authorities apply the 
name "thermodynamic efficiency" to the "thermal efficiency" as defined in para- 
graph 163. 

t This may be based on either indicated, brake or electrical horsepower. 



RECIPROCATING STEAM ENGINES 



365 



of perfection of the engine under the given conditions. This rate is 
called cylinder efficiency and may be expressed as 

2546 



^" W [{Hi - He) + {Pc - P2) XcUo -^ 778] 
See equations (138) and (132). 



(144) 



Example 28. Determine the cylinder efficiency of the two engines 
specified in paragraph 162. 
Superheated steam engine: 

j,r^ 2546 

8.5 1 [1332 - 980 + if| (2.0 - 0.5) 0.866 X 173.5] 
= 0.761 = 76 per cent. 

Saturated steam engine: 

2546 



E' 



12 [1176 - 935 + m (4 - 2) 0.808 X 90.5] 
= 0.808 = 80.8 per cent. 



Summing up the various efficiencies for the two cases analyzed in 
paragraphs 162 to 166: 



Saturated 
Steam Engine. 



Superheated 
Steam Engine 



Pressure, pounds per square inch, absolute: 

Initial 

Release 

Condenser 

Degree of superheat, deg. fahr 

Steam consumption, pounds per developed horsepower- 
hour: 

Actual engine 

Ideal engine, Rankine cycle, with incomplete expan- 
sion 

Ideal engine, Rankine cycle, with complete expansion 

Ideal engine, Carnot cycle 

Thermal efficiency, per cent: 

Actual engine 

Ideal engine, Rankine cycle, with incomplete expan- 
sion 

Ideal engine, Rankine cycle, with complete expansion 

Ideal engine, Carnot cycle 

Heat consumption, B.t.u., per developed horsepower- 
minute, Actual engine 

Ideal engine, Rankine cycle, with incomplete expan- 
sion 

Ideal engine, Rankine cycle, with complete expansion 

Ideal engine, Carnot cycle 

Rankine cycle ratio, per cent 

Cylinder efficiency, per cent 



150 
4 
2 

0.98=" 



12.00 

9.69 

9.16 

10.59 

19.6 

24.3 

25.8 
28.3 

216.4 

190.4 

174.8 

152.5 

76.3 

80.8 



200 

2 

0.5 
250 



8.50 

6.46 
6.00 



23.3 

30.7 
33.3 



181.9 

154.6 
138.4 

'70.'6' 
76.1 



* Quality. 

t If the steam consumption per i.hp-hour is used in this connection instead of 
the consumption per br.hp-hour this ratio becomes the indicated cylinder efficiency. 



366 STEAM POWER PLANT ENGINEERING 

167. Commercial Efficiencies. — There is no accepted standard for 
rating the commercial efficiency of an engine or turbine. The various 
measures used in this connection, such as pounds of standard coal per 
d.hp-hour, cents per horsepower per year and the like include the economy 
of the boiler and auxiliaries and are not a true indication of the per- 
formance of the engine alone. From a commercial standpoint it is im- 
portant to know the weight of coal required to develop a horsepower- 
hour, taking into consideration all of the losses of transmission and 
conversion, and a knowledge of the overall efficiency from switchboard 
to coal pile is of value in basing the cost of power, but these items 
are in reality measures of the plant economy and are of little value in 
comparing the performance of the prime mover. The various effi- 
ciencies under this heading are treated in Appendices B to G. 

168. Heat Losses in the Steam Engine. — The principal losses which 
tend to lower the efficiency of the steam engine and which prevent it 
from realizing the performance of the ideal engine are due to : 

(a) Cylinder condensation. 

(6) Leakage. 

(c) Clearance volume. 

(d) Incomplete expansion. 

(e) Wire drawing. 

(/) Friction of the mechanism. 

(g) Presence of moisture in the steam at admission. 

(h) Radiation, convection and minor losses. 

169. Cylinder Condensation. — The weight of steam apparently used 
per revolution, as determined from the indicator card, or the indicated 
steam consumption'^ (see paragraph 8, Appendix B), is considerably 
less than that actually supplied. The difference or missing quantity is 
due chiefly to cylinder condensation. This is by far the greatest loss in 
the steam engine with the exception of that inherent in the ideal en- 
gine. When steam is admitted to the cylinder a considerable portion 
of the heat is given up to the comparatively cool skin surface of the 
cylinder walls. If the steam is saturated at admission this heat ab- 
sorption causes condensation, or initial condensation as it is called; if 
superheated at admission the temperature is lowered to a correspond- 
ing point. After cut-off heat continues to be given up to the walls 
until the temperature of the steam falls below that of the skin surface, 
when the process is reversed and part of the heat is returned to the 
steam. With saturated steam the heat absorption causes condensation 
during expansion, and the heat rejected, reevaporation during expansion, 

* Also called the steam accounted for by the diagram or diagram steam. 



RECIPROCATING STEAM ENGINES 



367 



With superheated steam an equivalent heat exchange takes place. 
Unless the cylinder is of a compound series the heat absorbed from the 
cylinder walls during exhaust does no useful work and is lost. Cylinder 
condensation, measured as the proportion of the mixture present, 
varies with the size of the engine, speed, length of cut-off, valve design, 
temperature range, location of ports and port passages, jacketing, 
lagging, and other variables. It ranges from 16 to 30 per cent, and 
is occasionally as high as 50 per cent of the total weight of steam ad- 
mitted to the cylinder. Cylinder condensation and leakage cannot be 
conveniently separated and are ordinarily classified together. Fig. 191 



60 




O 
































\ 






























59 




s 


\, 










Condensation and Leakage 
for 
Simple Engines 

using 
Saturated Steamv 












\ 


o 

\ 














40 








\ 


\o 
































0^ 


V 






















30 










o 


o 


n 


^^^0 
































O 


^v^ 


< 














20 






















^ 


"^ 


' r 
































^ 


^ 


■-rrr 


O 


10 






















































E 


igine 


Tests 


Ban 


us, p. 


254. 



15 20 25 

Percentage of Cut Off 

Fig. 191. 



30 



35 



40 



shows the relation between cylinder condensation and leakage losses 
for various percentages of cut-ofT for simple high-speed non-condensing 
engines. 

Empirical formulas for calculating the extent of these losses, and 
which involve the various influencing factors, are unwieldy and only 
approximately accurate. One of the most satisfactory formulas of 
this class is that deducted by R. C. H. Heck, ''The Steam Engine and 
Turbine," 1911, p. 175. 

The various heat exchanges between working fluid and the cylinder 
walls, including cylinder condensation and leakage, are approximately 
determined by transferring the indicator diagram to the temperature 
entropy chart. (See paragraph 466.) 



368 



STEAM POWER PLANT ENGINEERING 



For use and application of the temperature-entropy diagram in 
engine tests consult Power, Dec, 1907, p. 834; Jan. 21, 1908, p. 96; 
Jan. 28, 1908, p. 145. 



131.3 
115.3 
96.3 
77.3 
58.1 




CRANK END 



133.0 

113. t 

95.0 

76.3 

57.1 



29.i 

19.5 

11.6 

0.0 



HEAD END 



131.3 
115.3 
96.3 
77.3 
58.1 





\ 




v 




V 








I 




^-.:^ ^ 







57.1 
38.9 
29.1 
19.5 
11.5 
0.0 



Fig. 192. Diagrams Taken from a 12-in, by 24-in. Corliss Engine. 

A comparatively simple method for approximating cylinder conden- 
sation and leakage losses is given by J. Paul Clayton, Proc. A.S.M.E., 
April, 1912, and consists in transferring the indicator diagram to loga- 
rithmic cross-section paper. By means of the logarithmic diagram 











_ 1 




cRNX 






.100 


L 


f- 














^ 90 












— 




































o< °" 


i 


















*>^ 




M ,„ 


i~ 






















^ '*> 


-i— 


















V^ 




*» rn 


\ 






















ft 60 


\ 


















\ 




m ciQ 


\ 


















\ 




s 

I in 


■\ 


\ 














D 


Y\ 




CR.\ 


H 














N 




S 


^ 


^ 
















^ 






\ 






















^ 


k 




















\ 


s 














1 






c 


"^ 


A 








_<^ 


V 


< 

10 










D< 


'^" 


B 









0.2 0.3 0.1 0.5 0.6 0,70.8 0.91.0 

Absolute Volume - Cu. Ft. 



Fig. 193. Logarithmic Diagrams Plotted from Fig. 192. 

Clayton found that (1) free from certain abnormal influences, expan- 
sion and compression take place in the cylinder substantially according 
to the law PV"" = C, (2) the value n bears a definite relation in any 



i 



RECIPROCATING STEAM ENGINES 



369 



given cylinder to the proportion of the total weight of steam mixture 
which was present as steam at cut-off, (3) the relation of the value n 
to the value x,.. (quality of steam at cut-off) for the same class of cyhnder 
as regards jacketing is practically independent of engine speed and of 
cylinder size, and (4) by means of the experimentally determined re- 
lation of Xn and n the actual steam consumption may be obtained 
from the indicator card to well within 4 per cent of the true value. 
The curves in Fig. 193 were plotted on logarithmic cross-section paper 
from the pressure-volume diagrams, shown in Fig. 192, and illustrate 




o Test with saturated Steam 
a Test with Superheated Steam 
• Center of gravity of all points 
Squation of condition X?=l. 258 n-0. 



0.900 0.950 1.000 1.050 1.100 1.150 1.200 1.250 
Value of n fx-om Expansion Curve 

Fig. 194. Relation of Quality and the Value of n. 

Mr. Clayton's method of analysis. The curves in Fig. 194 show 
the relation between quality Xc and exponent n or for a given set of 
conditions. See also paragraphs 470-3. 

Cylinder Condensation: Power, Jan. 3, 1911, p. 25; Jan. 30, 1912, p. 145; Nov. 5, 
1912, p. 664. 

170. Leakage of Steam. — The loss due to leakage is a variable 
factor depending upon the design and condition of the engine, and is 
greater with saturated than with superheated steam. The usual 
method of measuring leakage past the valves and piston while the 
engine is at rest is likely to give erroneous results, as demonstrated by 
Callender and Nicolson (Peabody, '^ Thermodynamics," p. 351) in tests 
made on a high-speed automatic balanced- valve engine and on a quad- 
ruple expansion engine with plain unbalanced slide valves. With the 
engines at rest they found that the leakage past valves and piston was 
insignificant, but when in operation the leakage from the steam chest 
into the exhaust was considerable. It was thought that a large propor- 



370 



STEAM POWER PLANT ENGINEERING 



tion of the leakage was probably in the form of water formed by con- 
densation of steam on the seat uncovered by the valve. 

According to the report of the Steam Engine Research Committee 
(Eng. Lond., March 24, 1905, p. 208), leakage through a plain slide 
valve is independent of the speed of the sliding surfaces, and directly 
proportional to the difference in pressure on the two sides; with well- 
fitted valves the leakage is never less than 4 per cent of the volume of 

steam entering the 
cylinders, and is often 
greater than 20 per 
cent. 

The various leakage 
losses may be approx- 
imated by transferring 
the indicator diagram 
to logarithmic cross- 
section paper. Fig. 195 
shows the apph cation 
of the logarithmic dia- 
gram to a specific case 
and illustrates this 
method of determining 
leakage losses. See 
paragraph 465. 

Leakage Past Piston 
Valves: Engr., Feb. 9, 1912. 



100 ri 
90 - 












= 




E 


p=m 


^±tt 






::= 




= 


= 


«0- 












= 






^- 


4:Tt 






— 




~ 


~ 


70 - 


.. 



















:t^ 






















if 






' 




— 








-^rr 






— 





— 


— 
























































i 


















\ 






























40 




V\ 


















V 
















\>, 


















N. 
















X 


^» 
















N 


















N 


















\ 










20 - 








\ 
















\ 




















\ 


V 














s 


^ 






10- 








^^ 




\ 


\ 


S 














) 


^ 


8- 


::: 






\ — 
1 — 




— 






-^2 








= 




t: 


- 




--- 






— 




— 1 






--^^ 


-^ — - 






— 




I — 


\— 


6 - 




































































^ 




















5 V 
















































4 






















■^ 




y 










































H 


































0.07 


^'' 


).l 


0. 


2 





3 





4 


0.5 0.6 


'■rs\ 


.0 


2 





3 





4.0 



Absolute Volume - Cu. Ft. 

Fig. 195. Diagram from a 14-in. by 35-in. Corliss 
Engine, Showing Leakage at Beginning and End of 
Expansion and Compression. 



171. Clearance Volume. — The portion of the cyhnder volume not 
swept through by the piston but which is nevertheless filled with steam 
when admission occurs is called the clearance volume. It is the space 
between the end of the piston when on dead center and the inside of 
the valves covering the ports. It varies from about 1 per cent of the 
piston displacement in very large engines with short steam passages 
to 10 per cent or more in small high-speed engines. 

The extent of surface in the clearance space greatly influences the 
amount of cyhnder condensation since the piston is moving slowly 
near the end of the cyhnder, and the time of exposure of the steam 
to these surfaces is comparatively long. The greater part of the cyl- 
inder condensation usually occurs in the clearance space, therefore 
the steam passages and clearance space should be designed so as to 
present a minimum amount of surface consistent with the proper 
cushioning volume for smooth operation. Theoretically if steam is 
compressed adiabatically to the initial pressure there is no loss due to 



RECIPROCATING STEAM ENGINES 



371 



clearance but in practice compression carried to initial pressure does 
not necessarily improve the economy. For a constant time element 
the shorter the cut-off the greater will be the ratio of the weight of 
cushion steam to that of the steam supplied and hence the greater the 
loss. In large, slow-moving engines the loss due to clearance may be 
greater than that in high-speed, short-stroke engines because of the 
longer time of exposure to the clearance surface. 

The ratio of expansion is decreased by clearance; for example, an 
engine cutting off at one-fifth, neglecting clearance, has an apparent 
ratio of expansion of 5, but if the clearance volume is 10 per cent the 
actual ratio is only 3.66. One of the few recorded tests relative to the 
influence of clearance on the economy of a high-speed engine was con- 
ducted on a 14-in. by 15-in. Allfree engine. (Power, May, 1901.) 
With a clearance volume of 2.2 per cent, initial pressure 105 pounds 
gauge, and 172 r.p.m., the best performance was 23.7 pounds of dry 
steam per i.hp-hour. With the same steam pressure and speed but 
with clearance volume increased to 6 per cent by the use of a shorter 
piston, the best performance was 28.3 pounds per i.hp-hour. In both 
cases the compression was carried up to admission pressure. 

Independent tests made by Prof. Boulvin and by A. H. Klemperer 
on single-cylinder CorUss engines gave a minimum water rate when the 
clearance volume was approximately one-half the compression volume. 
See end of paragraph 172. 

Engine Clearance and Compression: Power, July 5, 1910, Dec. 27, 1910; Sibley 
Journal, Dec, 1910. 

173. Loss Due to Incomplete Expansion and Compression. — Theo- 
retically the loss due to incomplete expansion is considerable. For 
example, the theoretical steam consumption of a perfect engine (Ran- 
kine cycle) expanding from 120 pounds absolute to a condenser pres- 
sure of 2 pounds absolute is 9.6 pounds per horsepower-hour. If the 
expansion were carried to only 5 pounds absolute, the exhaust pressure 
remaining the same, the steam consumption would be increased to 
11.8 pounds per horsepower-hour, a difference of 22 per cent for an 
increase in terminal pressure of only 3 pounds per square inch. The 
theoretical water rates for various terminal pressures are given below. 



Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


1 

1.5 

2 

2.5 


8.5 
9.1 
9.6 
10 


3 
4 
5 
6 


10.4 
11.1 
11.8 
12.3 



372 



STEAM POWER PLANT ENGINEERING 



In actual engines expansion is seldom complete, since it would neces- 
sitate increased bulk and weight of engine, and the work done by the 
steam in the last stages would not compensate for the increased cost. 

In single-cylinder engines maximum economy is effected when the 
terminal pressure is considerably above that of the exhaust, since the 
gain due to complete expansion is more than offset by the increased 
cyhnder condensation. This is true to a certain extent in all engines 
irrespective of the number of cyHnders. Tests by G. H. Barrus 
(''Engine Tests," 1900) to determine the terminal pressures effecting 
maximum economy for various types of engine gave results as follows: 



Simple slide-valve engines, non-condensing 
Simple slide-valve engines, condensing .... 
Simple Corliss engines, non-condensing. . . . 

Simple Corliss engines, condensing 

Compound engines, non-condensing 

Compound engines, condensing 



Terminal Pressure, 


Pounds Absolute. 


30 to 40 


25 to 30 


20 to 25 


15 to 18 


18 to 22 


3 to 5 



In high-speed engines a certain amount of compression is desirable 
for its cushioning effect; outside this mechanical feature compres- 
sion may or may not be of benefit to the engine, as will be seen from 
the results of tests stated below. Zeuner in his treatise on theoreti- 
cal thermodynamics proves deductively that in an engine with a large 
clearance volume the loss due to clearance is completely eliminated if 
the compression is carried up to admission pressure, a conclusion which 
tests by Jacobus, Carpenter, and others fail to confirm. A series of 
tests by Professor Jacobus (Trans. A.S.M.E., 15-918) on a 10-in. by 
11-in. high-speed automatic engine at Stevens Institute show decreasing 
economy with increase of compression, the initial pressure, cut-off, and 
release remaining constant. The results were as follows: 



Proportion of initial pressure 

steam is compressed 

Steam, pounds per i.hp-hour. . . 



up to which the 




Tests by Carpenter (Trans. A.S.M.E., 16-957) on the high-pressure 
cylinders of the CorHss engine at Sibley College gave: 



Compression, per cent 

Brake horsepower 

Steam, pounds per br.hp-hour. 




RECIPROCATING STEAM ENGINES 



373 




Fig. 196. 3000-hp. Sulzer Engine Designed for Highly Superheated Steam. 



374 



STEAM POWER PLANT ENGINEERING 



Tests made by A. H. Klemperer on a 7.1-in. by 17.7-m. Corliss 
engine, at Dresden, gave decreasing steam consumption for increase 
in compression up to a compression of about twice the clearance vol- 
ume, beyond which the water rate increased with the increase in com- 
pression. (Zeit. d. Ver. deut. Ingr., Vol. I, 1905, p. 797.) 

Tests made by Prof. Boulvin on a 9.8-in. by 19.7-in. Corliss engine at 
University of Ghent gave results agreeing with those of Klemperer. 
(Revue de Mecanique, 1907, Vol. XX, p. 109.) 



6.8 

6.7^ 

6.6 

6.5 

6.4 

6.3 

6.2 

6.1 

6.0 

5.9 

5.8' 

5.7 

5.6 

5.5 














. 








18 
47 
46 
45 

w 














100 Lb. Gauge 


^^ 


\ 
















^ 




\ 














^ 










'^ 






J 


-^^ 












^ 






^^ 


^v 
















^ 


^^\ 












i*} 






J 




\^ 












41 

10 






r^ 




'(-V 


^«>^ 










J 


^ 










^^ 








39 
38 
37 


^ 




Influence of Back 

Pressure on the Economy of 

an 8 X 10 Automatic High Speed 

Non Condensing Engine 


^\^ 
























^^ 




















'35 



2 4 6 8 10 12 14 10 18 20 

Back Pressure.Lb. Per Sq.In.Gauge 

Fig. 197. 

Fig. 197 shows the influence of increasing back pressure on the 
economy of an 8-in. by 10-in. automatic high-speed engine at the 
Armour Institute of Technology. 
, The Effeci zj Compression: Power, Oct. 27, 1914, p. 595. 

173. Loss Due to Wire Drawing. — Wire drawing, or the drop in 
pressure due to the resistances of the ports and passages, has the effect 
of reducing the output and the economy of the engine to some extent, 
since the pressure within the cylinder is less than that at the throttle 
during admission and greater than discharge pressure at exhaust. 
The steam may be dried to a small extent during admission, but be- 
cause of the drop in pressure the heat availability is reduced. The loss 
in available heat may be calculated as shown in paragraph 456. In 
single-valve engines the effects of wire drawing are decidedly marked 
and the true points of cut-off and release are sometimes difficult to 
locate on the indicator card. In engines of the Corliss, poppet; or 
gridiron-valve type the effects are hardly noticeable. 



RECIPROCATING STEAM ENGINES 



375 



174. Loss Due to Friction of tlie Mechanism. — The difference between 
the indicated horsepower and that actually developed is the power 
consumed in overcoming friction, and varies from 4 to 20 per cent of 
the indicated power, depending upon the type and condition of the 
engine. Engine friction may be divided into (1) initial or no-load 
friction and (2) load friction. The stuffing-box and piston-ring friction 
is practically independent of the load, while that of the guides, bear- 
ings, and the Uke increases with the load. In Fig. 198, curve A gives 
the relation between the frictions for a four-slide-valve horizontal 
cross compound engine, and B that for a simple non-condensing Corhss. 




100 125 150 115 200 225 
Developed Horse-Power 



275 300 



Fig. 198. Typical Curves of Steam Engine Friction. 



(Peabody's ''Thermodynamics," pp. 433 and 437.) Curve C is plotted 
from the tests of a Reeves vertical cross compound condensing engine 
(Engineering Record, July 1, 1905, p. 24), and D from the test of an 
Ames simple high-speed non-condensing engine. (Engineering Record, 
Vol. 27, p. 225.) A large number of recorded tests show less friction at 
full load than at no load, but this is probably due to error or to varia- 
tions in lubrication. With first-class lubrication it is usually sufficiently 
accurate to assume the friction to be constant and equal to the initial 
friction at zero load. The distribution of the frictional losses in a 
number of engines is given in Table 73. 

175. Moisture. — The presence of moisture in the steam pipe is due 
to condensation caused ))y radiation or to priming at the boiler. Unless 
removed by some separating device between boiler and engine the 
amount of moisture entering the cylinder may be from 1 to 5 per cent 
of the total weight of steam, and the work done per pound of fluid is 
correspondingly reduced. This loss should not be charged against 
the engine, however, and its performance should be reckoned on the 



376 



STEAM POWER PLANT ENGINEERING 



dry steam basis. Experiments reported by Professor R. C. Carpenter 
(Trans. A.S.M.E., 15-438) in which water in varying quantities was 
introduced into the steam pipe, causing the quaUty of the steam to 
range from 99 per cent to 57 per cent, showed that the consumption 
of dry steam per i.hp-hour was practically constant, the water acting 
as an inert quantity. An efficient separator will remove practically 
all the entrained water. 

TABLE 73. 
DISTRIBUTION OF FRICTION IN SOME DIRECT-ACTING STEAM ENGINES. 

(Thurston.)* 





Percentage of Total Engine Friction. 


Parts of Engines where Friction 
is Measured. 


" Straight 
Line " 
Balanced 
Valve. 


" straight 
Line " 
Unbalanced 
Valve. 


Traction 
Engine 
Locomotive 
Valve Gear. 


Automatic 

Balanced 

Valve. 


Condensing 
Engine 
Balanced 
Valve. 


Main bearings 


47.0 


35.4 


35.0 


41.6 


46.0 


Piston and piston rod 


32.9 


25.0 


21.0 


49.1 








Crank pin 


6.8 


5.1 


13.0 


21.8 


Crosshead and wrist pin 


5.4 


4.1 




Valve and valve rod 


2.5 


26.4 


22.0 


9.3 


21.0 


Eccentric strap 


5.4 


4.0 






Link and eccentric 






9.0 
















Air pump 










. 12.0 
















100.0 


100.0 


100.0 


100.0 


100.0 



Friction and Lost Work in Machinery," p. 13. 



176. Radiation and Minor Losses. — The radiation and conduction of 
heat from the cylinder, piston rod and valve stem has the effect of in- 
creasing the cylinder condensation. In jacket engines this loss may be 
approximated by the quantity of steam condensed in the jacket when 
the engine is not running. In unjacketed engines the loss is practically 
undeterminable since the heat exchange between cylinder walls and 
the steam is exceedingly complex. The heat loss due to radiation, 
measured in terms of the total heat supphed, varies from 0.3 per cent 
in very large units with efficiently lagged cylinders and steam chests 
to approximately 2 per cent in small engines as ordinarily insulated. 

177. Heat Lost in tlie Exhaust. — Most of the heat supphed to the 
engine is rejected to the exhaust; this varies from 70 per cent in the 
most economical type of prime mover to 95 per cent in the poorer 



RECIPROCATING STEAM ENGINES 377 

types. If the exhaust steam is used for heating purposes the heat 
chargeable to power is the difference between the heat supphed and 
that utiHzed from the exhaust. In passing through a prime mover 
heat is abstracted from the steam by: 

(1) Conversion of part of the heat into mechanical energy. 

(2) Loss through radiation. 

If w represents the water rate or steam consumption per indicated 

2546 
horsepower-hour or the equivalent, then — — = B.t.u. utilized per 

hour from each pound of steam in producing one indicated horsepower. 
Considering Hi as the initial heat content, B.t.u. per pound above 
32 degs. fahr., and Hr as the loss due to radiation, the heat content 
H2 per pound of exhaust will be 

H, = H,-Hr- ?^- (145) 

w 

As previously stated the heat loss due to radiation in terms of the 
total heat supplied varies from 0.3 per cent in very large units with 
efficiently lagged cylinders and steam chests to approximately 2 per 
cent in small engines of 25 horsepower rated capacity. An average 
value of one per cent may be assumed for most practical purposes. 

If the exhaust contains moisture as is usually the case, we have 

H2 = X2r2 + 52, (146) 

in which 

X2 = quality of the exhaust, 

r2 = latent heat corresponding to exhaust pressure, 

52 = heat of the liquid at exhaust pressure. 

Combining equations (145) and (146) and reducing 

x, = ^. (147) 

T2 

If the exhaust is superheated 

H2 = r2 + q2 + Cmk\ (148) 

in which 

Cm = mean specific heat of the superheated steam at exhaust pressure, 
^' = degree of superheat of the exhaust steam, deg. fahr. 

Assuming that the moisture in the exhaust is rejected to waste, 
the heat available per pound of exhaust steam at the exhaust nozzle 
is X2r2 + Q2 and the net heat chargeable to power is 

w [Hi — {X2r2 + 52)] B.t.u. per i.hp-hr. (149) 



378 



STEAM POWER PLANT ENGINEERING 




Fig. 199. 7500-kw. Vertical-horizontal Double-compound Engine as Installed at 
the 59th Street Station of the Interborough Rapid Transit Co. (Manhattan Type.) 

All of this heat is not available for commercial heating purposes 
because of the condensation losses in the exhaust main. The extent 
of the latter depends upon the size and length of main, rate of flow 
and efficiency of the pipe covering. Representing this loss, per pound 
of steam, by H^, the total heat chargeable to power is 

w [Hi — {x^Ti + 52) + H^ B.t.u. per i.hp-hr., (150) 

and the equivalent water rate for power only is 



w \Hi — {X2r2 + g2) + HJ 
H 



lb. per i.hp-hr., 

£1 

in which 
H = net heat supplied to the engine, B.t.u. per pound. 



(151) 



RECIPROCATING STEAM ENGINES 379 

Very little information is available relative to the quality of exhaust 
as determined by actual test but such as has been pubUshed is in accord 
with the results calculated from equation (147). 

Moisture in Exhaust Steam. — Trans. Am. Soc. Heat & Vent. Engrs., Vol. 21, 1915, 
p. 85; Trans. A.S.M.E., Vol. 32, p. 331. 

Example 29. A 23-inch by 16-inch simple engine, direct connected 
to a 200-kilowatt generator installed at the Armour Institute of Tech- 
nology uses 35 pounds of steam per indicated horsepower-hour at full 
load, initial pressure 115 pounds absolute, back pressure 17 pounds 
absolute, initial quality 98 per cent. 

Calculate the quality of the exhaust, assuming a radiation loss of 
one per cent. 

From steam tables Hi = xr -\- q 

= 0.98 X 879.8 -I- 309 
= 1171+, 
r2 = 965.6, 
§2 = 187.5. 

By assumption, Hr = 0.01 X 1171 = 11.7. 
Substituting these values in equation (147) 

2^46 
1171 - 11.71 - 187.5 - — - 

X2 = TTTrTT '' — = 0.933 or 93.3 per cent. 

965.6 

(Actual calorimeter tests gave a quality of 92.5 per cent, indicating 
a somewhat larger radiation loss than the assumed value of one per 
cent.) 

Total heat chargeable to power (equation 150). 

35 [1171 - (901 + 187.5) -f- H,] = 2918 -j- 35 H, B.t.u. per i.hp-hr. 

Hx varies within such wide limits that general assumptions are apt 
to lead to serious error. Where specific figures are not available it is 
customary to allow 2 per cent of the heat value of the exhaust as the 
extent of this loss. With this assumption we have as the heat charge- 
able to power 

2918 -h 35 X 0.02 (901 -h 187.5) = 3680 B.t.u. per i.hp-hr. 

Assuming that the condensation from the heating system, including 
that exhausted from the engine, is returned to the boiler at a tempera- 
ture of 192 deg. fahr., the net heat supplied per pound of steam is 

H = 1171 - (192 - 32) 
= 1011 B.t.u. 

And the equivalent water rate for power only is 
3680 



1011 



3.63 lb. per i.hp-hr. 



The low fuel consumption for power when the exhaust steam is used 
for heating purposes is at once apparent. 



380 



STEAM POWER PLANT ENGINEERING 



178. Methods of Increasing Economy. — Various methods have been 
adopted for bettering the economy of piston engines; among them may 
be mentioned : 

(a) Increasing boiler pressure. 
(6) Increasing rotative speed. 

(c) Decreasing back pressure by condensing. 

(d) Superheating. 

(e) Use of steam jackets. 
(/) Reheating receivers. 
{g) Compounding, 

(h) Use of uniflow or straight-flow cylinders. 
(i) Use of binary fluids. 

179. Increasing Boiler Pressure. — A glance at Table 74 will show 
that increase in initial pressure, other conditions remaining the same, 
results in increased theoretical eflficiency. This increase is so marked 
that engineers are considering the possibilities of employing pressures 
far above any now in use. There is no question but that working 
pressures as high as 600 pounds per sq. in. abs. will necessitate radical 
departure from the present type of boiler and may involve costs which 
are prohibitive, but the present tendency is toward the higher pressures. 
The design of engines for high pressures is not a difficult one since 
pressures as high as 800 pounds per sq. in. abs, are used successfully 
in Diesel engines. Several steam turbine plants are now under con- 
struction in which initial pressures of 350 pounds gauge and tempera- 
tures of 650 deg. fahr. are to be used, but until actual operating data 
are available no conclusions can be drawn as to the ultimate conmier- 
cial economy effected by this practice. With the ordinary type of 



TABLE 74. 

THEORETICAL EFFICIENCY. 

Rankine Cycle. 

Initial Temperature Constant (600 Deg. Fahr.). 





Superheat, Deg. Fahr. 


Efficiency, Per Cent. 


Initial Pressure, 
Lb. per Sq. In. Abs. 


Condensing Back Pressure 
i Lb. Per Sq. In. Abs. 


Non-condensing Back 

Pressure 14.7 Lb. 

Per Sq. In. 


1574 
600 
500 
400 
300 
200 
100 



113.4 
132.7 
155.2 
182.5 
218.1 
272.2 


40.3 
37.3 
36.7 
36.1 
34.5 
32.9 
29.8 


30.6 
26.0 
25.0 
23.7 
22.0 
19.7 
15.4 



RECIPROCATING STEAM ENGINES 



381 



double-flow engine the heat economy increases with the pressure up 
to the point where increased condensation losses and leakage neutralize 
the theoretical gain. This point of maximum efficiency varies with 
the size and type of engine and the grade of workmanship. 

Fig. 200 shows the results of tests made at the Armour Institute 
of Technology on an 8-in. by 10-in. automatic high-speed piston- valve 
engine. A marked gain will be noted up to a pressure of 115 lb. per 



6.2 
^ 6.1 
S 6 


Nv 
















. 


A 


X, 










,y^ 


^ 


"^ 




\ 


h 






/ 








.§5.9 

e 

|5.8 

^5.7 
56 






*^^ 


X 


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PX 


^s; 


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. 


5.5 


• 


/■ 












y 


'■^^ X 



46 



45 



43 a, 



110 



115 



120 



75 80 85 90 95 100 105 

Initial Gauge Pressure, Lb.perSq. In. 

Fig. 200. Influence of Initial Pressure on the Economy of a Small, High-speed, 
Non-condensing Engine. 

sq. in. gauge beyond which the gain is very small. The following 

figures show the increase in economy with increased boiler pressure 

in a consohdated locomotive engine. (Bulletin No. 26, University of 

Illinois, Experimental Station.) 

Boilerpressure, lb. persq. in... 120 140 160 180 200 220 240 
Steam per i.hp-hr., lb 29.1 27.7 26.6 26.0 25.5 25.1 24.7 

A small Willans engine, non-condensing, gave results as follows: 

Initialpressure, lb. persq. in... 36.3 51.0 74.0 85.0 97.0 110.0 122.0 
Steam per i.hp-hr., lb 42.8 36.0 32.6 29.7 26.9 27.8 26.0 

The uniflow engine offers possibiUties for high pressures which are 

very promising but the art, at least in America, is still in an experimental 

stage. Tests on 100 hp. condensing engines by Lentz gave the following 

results : 

Initial steam pressure, lb. per sq. in. abs 235 461 

Initial temperature, deg. fahr 923 1018 

Steam consumption, lb. per i.hp-hr 6.52 5.67 

Heat consumption, B.t.u. per i.hp-minute 162 144 



382 



STEAM POWER PLANT ENGINEERING 



The range of pressures sanctioned by modern practice for different 
types of engines is as follows: 



Type of Engine. 


Range in Pressure 
(Gauge). 


Average. 


Simple slow-speed (standard type) 


60-120 
70-125 
125-225 
100-180 
100-180 
125-200 
140-250 
175-300 


90 


Simple high-speed (standard type) 


100 


Simple, uniflow (condensing) 


175 


Compound high-speed, non-condensing 


150 


Compound high-speed, condensing 


150 


Compound slow-speed, condensing 


170 


Triple expansion, condensing 


200 


Quadruple expansion, condensing 


250 



Higher Steam Pressures: R. Cramer, Trans. A.S.M.E., Vol. 37, 1915, p. 597. 
High Pressure Steam for Superheating: Power, Dec. 28, 1915, p. 892. 

180. Increasing Rotative Speed. — High rotative speed does not 
necessarily mean high piston speed. An 8-in. by 10-in. engine running 
at 300 r.p.m. has a piston speed of only 500 feet per minute, whereas a 
36-in. by 72-in. Corhss running at 60 r.p.m. has a piston speed of 720 
feet per minute. The classification ''high speed" and "low speed" 
refers to rotative speed only, the former above and the latter below, 
say 150 r.p.m. 

On account of the reduction of thermodynamic wastes, a high-speed 
engine should give theoretically a higher efficiency than the same 
engine at a lower speed, all other conditions being the same. The effect 
of speed upon economy is decidedly marked in engines and pumps 
taking steam full stroke. For example, tests of a 12-in. by Tj-in. by 
12-in. simplex direct-acting steam pump at Armour Institute of Tech- 
nology showed a steam consumption of 300 pounds per i.hp-hcur at 
10 strokes per minute, and only 99 pounds at 100 strokes per minute. 
(See Figs. 381 and 382.) 

Tests of engines using steam expansively, however, do not furnish 
conclusive evidence on this point, some showing a decided gain (Pea- 
body, ''Thermodynamics," p. 425), others httle or no gain (Barrus, 
"Engine Tests," p. 260). For example, a small Willans engine showed 
an increase in economy of 20 per cent in increasing the rotative speed 
from 111 to 408 r.p.m. (Peabody, "Thermodynamics," p. 402), whereas 
the compound locomotive at the Louisiana Purchase Exposition showed 
a loss in economy for the higher speeds (Publication by the Pennsyl- 
vania Railroad Company). On the other hand, a comparison of the 
performances of high- and low-speed Corliss engines shows little differ- 
ence in economy, and a general comparison between high- and low- 
speed engines furnishes little information, since nearly all high-speed 



RECIPROCATING STEAM ENGINES 383 

engines are of a different class from the low-speed ones. High-speed 
engines are comparatively small in size, require larger clearance volume, 
and are usually fitted with a single valve. Rotative speed is limited V)y 
design, material, workmanship, and cost of subsequent maintenance. 
Speeds of 400 r.p.m. and more are not unusual with single-acting 
engines, whereas 300 r.p.m. is about the limit for double-acting machines 
with strokes over 12 inches in length. A comparison of tests of high- 
speed and low-speed engines in this country, irrespective of design and 
construction, shows the former to be less economical than the latter in 
most cases. In Europe high-speed engines are developed to a high 
degree of efficiency, and their performances are comparable with the 
best grade of low-speed engines. 

High-speed engines as a class have the advantage of being more 
compact for a given power, are simple in construction and relatively 
low in first cost; on the other hand, they are subject to comparatively 
rapid depreciation, excessive vibration, and are less economical in 
steam consumption. 

181. Decreasing Back Pressure by Condensing. — The effect of the 
condenser upon the power and economy of engines is indicated in Table 
75. The curves in Figs. 201 and 202 were plotted from tests made by 
Professor R. L. Weighton on a 7, lOj, 15| by 18-in. triple-expansion 
engine at Durham College of Science, Newcastle-on-Tyne. The 
straight line shows how the mean effective pressure would vary with 
the degree of vacuum if the power increased directly with the reduction 
in back pressure. The curved line shows the actual m.e.p., which 
increases almost along the theoretical line up to a 10-inch vacuum, 
from which point on the increase is less marked. At 26 inches the 
actual m.e.p. reaches an apparent maximum. These figures are not 
applicable to all engines but give a good idea of the limitation of the 
vacuum with the average type of reciprocating engine with restricted 
exhaust port openings. With specially designed ports and passages 
of large cross-sectional area the piston engine shows increase in steam 
economy up to the highest vacuum carried in the condenser. (See 
Power, Jan. 16, 1912, p. 72.) 

The gain in steam consumption due to the condenser does not indi- 
cate a corresponding gain in heat consumption. For example, Engine 
No. 2, Table 75, shows an apparent gain in steam consumption, due 
to condensing, of 12.5 per cent, the temperature of the feed water 
returned to the boiler being 120 deg. fahr. With a suitable heater the 
exhaust of the non-condensing engine would be capable of heating the 
feed water to 210 deg. fahr. The non-condensing engine should there- 
fore be credited with 210 — 120 or 90 heat units per pound of steam 



384 



STEAM POWER PLANT ENGINEERING 



43 



41 



-: 40 



37 



■B 36 



wi 


34 










1 


33 






^ 


32 


« 




Ph" 


31 


H 




n 


30( 



HB 



























// 


/ 


























y 


/^ 


























/ 


y 


























y 


/' 


























..^> 


K^-^S 


<) 




p 
















>-^^ 


(^ 




o 






















.1^ 


y^ 
























.s 


^ 


























A 


^ 


























/ 


X 


























/ 


/^ 


























y 




























/ 








Inc 


tease 
Trip 


inPc 


wer ] 
pansi 


Due to Vac 
on Engine 


uum 










' 

























































































10 12 14 16 18 20 22 24 
Vacuum in laches of Mercury 

Fig. 201. 



26 28 



19 



I 



•17 



16 



t5 

































380 
370 
360 
350 


\ 






I 


icrej 


rse in Economy due 
rripVe Expansi|on E 


to V 
igine 


ICUU 

1 


n 








\, 


\ 




























s 


\ 


\ 






























\ 


>v 


s^^ 






















OJA 








\ 


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^ 


330 










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. ^". 


y 


S'^O 














•^ 


^ 


^ar 


oi. 


pet 


^ 


^ 


























^< 


^: 


^I., 


2l?? 


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qno 
























' 






D 


290 



8 10 12 14 16 18 20 22 ^ 
Vacuum in Inches of Mercury 

Fig, 202. 



28 30 



RECIPROCATING STEAM ENGINES 



385 



used, or, in round numbers, 9 per cent. The difference between 12.5 
per cent and 9 per cent, or 3.5 per cent, represents the net gain in favor 
of condensing, provided the power necessary to create the vacuum is 
ignored. Actually, the steam consumption of the condenser pumps 
might be equal to or greater than 3.5 per cent of the steam generated 
and the net gain becomes zero or even negative. Referring to Fig. 
202, plotted from tests of the 7, 10|, 15^ by 18-in. triple-expansion 
engine mentioned above, the curves show the feed-water consumption 
per i.hp-hour and the heat units consumed per brake horsepower per 
minute measured above the hot- well temperature. The engine effi- 
ciency, based upon the water consumption, increases as the vacuum 
increases, reaching a maximum between 26 and 28 inches, whereas the 
heat-unit curve gives the maximum between 20 and 21 inches. Be- 
tween 22 and 28 inches the heat-unit curve shows a rapid falling off 
in economy. Tests of the 5500-horsepower engine at the New York 
Edison Company's Waterside Station showed that increasing the 
vacuum from 25.3 to 27.3 inches decreased the water rate only 0.06 
pound per i.hp-hour. (Power, July, 1904, p. 424.) The results are 
illustrated in Fig. 220. In most cases, and particularly with large 
compound engines, the net gain due to condensing is considerable, but 
the feed-water temperatures and power consumed by the auxiliaries 
should be taken into account. 

TABLE 75. 

EXAMPLES OF THE EFFECT OF CONDENSING ON THE ECONOMY OF SMALL 
RECIPROCATING ENGINES. 





Non-Condensing. 


Condensing. 


Increase Due 
to Condensing. 


if 


3 Si 

"S 


ll 
IS 


Ills 

5 S 3 1 


1. 

o 1 
'S 


>- C «3 ^ 

So ^^ 


% s. 
SI 


5.2 i2W 


ll 

e 

hi 


ll 

a 


1 


147 


54.7 


19.2 


149 


1.6 


83.4 


14.8 


52.5 


25 


2 


148 


540 


19.3 


147 


4 




16.9 


* 


12.5 


3 


126 


83 


23.8 


130 


7.4 


116 


19.1 


39.8 


19.7 


4 


67.6 


209 


28.9 


67 


4.5 


213 


22 


1.9 


23.5 


5 


103.8 


177.5 


22.1 


103.8 


1.2 


155 


16.5 


* 


25.1 


6 


114 


160 


31 


114 




168 


27 


2 


12.9 


7 


96 


120 


23.9 


96 


4 


145 


19.4 


20.8 


18.8 


8 


118 


267 


23.24 


119 


4.2 


276.9 


16 


3.7 


31 


9 


75.9 


310 


25.6 


79 


6.4 


336 


20.5 


8.7 


19.9 


10 


62.5 


451 


30.1 


63.6 


7.8 


444 


23 


* 


23.6 


11 


186.7 


40.4 


18.7 


184.6 


1.6 


29.8 


12.7 


* 


32 



Cut-off changed for best economy. 



386 



STEAM POWER PLANT ENGINEERING 



182. Superheating. — The theoretical gain due to the use of super- 
heat is comparatively small as will be seen from Table 76. Consider- 
ing the additional expense of equipment and maintenance of super- 
heating apparatus the ultimate gain would appear to be a negative 
quantity. Practically, however, the heat economy of the piston 
engine is greatly increased by superheating. This apparent anomaly 
is due to the fact that the theoretical engine is assumed to operate 
in a non-conducting cycle and no condensation takes place except in 
doing work, whereas, in the actual mechanism the cylinder is far from 
being non-conducting and considerable initial condensation takes place. 
The reduction of cylinder condensation due to the use of superheated 
steam is the principal reason for the marked gain in economy of the 
actual engine. The greater the cylinder condensation the larger is the 
saving possible. As a rough approximation the steam consumption 
is reduced about 1 per cent for every 10 deg. fahr. increase in super- 
heat but the actual value depends upon the type and size of engine 
and the initial condition of the steam. In American practice super- 
heat corresponding to a total steam temperature of 650 to 700 deg. 
fahr. appears to be the limit of commercial economy but in Europe 
temperatures as high as 900 deg. fahr. have been employed with 
apparent ultimate economy. 



TABLE 76. 

THEORETICAL EFFICIENCIES AND WATER RATES. 

Rankine Cycle — Superheated Steam. 

Initial Pressure 200 Lb. Per Sq. In. Abs. 





Effi 


ciency. 


Water Rate. 


Superheat, 










Deg. Fahr. 












Condensing.* 


Non-condensing. 


Condensing.* 


Non-condensing. 





31.88 


18.60 


6.94 


13.44 


50 


32.03 


18.71 


6.72 


12.96 


100 


32.24 


18.92 


6.52 


12.49 


150 


32.49 


19.18 


6.34 


12.03 


200 


32.77 


19.51 


6.16 


11.57 


250 


33.09 


19.89 


5.98 


11.10 


350 


33.81 


20.76 


5.67 


10.20 


400 


34.20 


21.25 


5.48 


9.77 


500 


35.04 


22.12 


5.16 


9.00 



Absolute back pressure 0.5 lb. per sq. in. 



Table 83 gives test results for several different types of engine em- 
ploying superheated steam. These figures may be^ compared with the 
performances of engines using saturated steam as given in Tables 81 and 
82. A decided gain in economy is shown in favor of superheat for single- 



RECIPROCATING STEAM ENGINES 



387 




388 



STEAM POWER PLANT ENGINEERING 



Indicated Horse Power 
40 60 80 100 120 140 160 180 200 220 



280 300 

7.000 




Per Cent, Load on Engine 
Fig. 204. Influence of Superheat on the Water Rate of a 16-inch by 22-inch 
Ideal Corliss Engine. 




200 

Superheat, Deg. F. 
Fig. 205. Effect of Superheat on Steam Consumption. 



RECIPROCATING STEAM ENGINES 



389 



cylinder engines. With compound engines the advantage is not so 
apparent, while triple-expansion engines show the least gain. Tables 
83 to 85 show the effect of superheating on simple, compound and triple- 
expansion engines. Some idea of the wonderful fuel economy effected 
in Europe with the use of highly superheated steam in connection with 
the so-called locomobile is gainedfrom the results shown in Table 80. 
This type of engine has not yet been introduced to any extent in this 
country but it is only a matter of time when the cost of coal will advance 
to such a point as to preclude all but the more economical types of prime 
movers. 

As far as steam consumption is concerned, all engines show greater 
economy with superheated than with saturated steam, but the thermal 



20 




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\ 
















•i 18 


1 




X 
V 


\^ 
















\ 




















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w 

n 16 


\ 


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^ted stp 


im. 




\ 


\ 












Supe 


rheat.-SO 


F. 






|'14 


\ 


\ 




^ 


"^^ 







<f 


<• -IOC 


°F. 


\ 


\^ 






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17 


'« -150 


'F. 


rj 








a 


\ 


v^ 






^"^ 






a 


" -200 


'F. 


55 12 












V, 






" 






,, 


" -250 


'F. 






r. 


" -300 


'F. 


















,, 


" -350 


K 


10 



















40 60 

Per Cent of Rated. Load. 



80 



100 



Fig. 206. Effect of Superheat on Steam Consumption. 

gain is not so marked, and when the economy is measured in dollars 
and cents per developed horsepower, taking all things into consideration 
the gain is still further reduced and in some cases completely neutralized. 
First cost, maintenance, and disposition of the exhaust must all be 
considered in determining the ultimate commercial gain due to the use 
of superheated steam. 

Fig. 205 gives the results of a series of tests made on a number of 
Belliss & Morcom engines using superheated steam. (Pro. Inst, of 
Mech. Engrs., March, 1905, p. 302.) The engines were from 200 to 
1500 kilowatts capacity and were tested at full load. It is noticeable 



390 



STEAM POWER PLANT ENGINEERING 



that the curves all converge to a single point and will meet at about 
400 deg. fahr. The results show that if sufficient superheat is put into 
the steam all engines of whatever size are equally economical. 

These curves though strictly applicable to the specific cases cited 
are more or less general and represent the influence of superheat on all 
types of piston engines. 







74.3 

1 






93.8 
1 


Degrees of Superheat, Fahr. 

106.8 115.97 

1 1 




133.5 


24 

C 

w 

i 20 

i 18 

14 




1 
1 




(A) 
(B) 


16|s22 
19 |x 21 


Corliss, 
Poppet 


200 r.p 
Valve i 


00 li.p.i 


b. initi 
I. 131 11 


il press 
. initia 


pfessu 


■e 








M 






1 
1 
1 






1 
1 

1 






1 

1 
1 










In 


\ 

^ 




1 

1 






1 
1 

1 






1 
1 
1 










1 


\^ 


^^--^ 


1 

1 




A 


1 
1 

— t— 








1 





— 1 






1 

1 






1 


^ 


B 


1 
1 






1 
1 




1 






1 
1 
1 






1 

1 
1 






1 
1 
1 






1 
'SI 










1 
1 

1 






1 
1 
1 






1 
1 






H-ll 









Fig. 207. 



60 



100 



140 160 180 200 220 
Indicated Horse Power 



240 260 280 



Comparative Water Rates of a Corliss Four-valve and a Poppet Four- 
Valve, High-speed Engine. 



183. Jackets. — If the walls of the cylinder are made double and 
the space between is filled with live steam under boiler pressure, the 
cylinder is said to be steam jacketed. The function of the jacket is 
to reduce initial condensation by maintaining the temperature of the 
internal walls as nearly as possible equal to that of the entering steam. 
The heat given up by the jacket steam, and the resulting condensa- 
tion, is usually a smaller loss than would otherwise result from cylinder 
condensation. However, tests of numerous engines with and without 
steam jackets do not agree as to the conditions under which their use 
is profitable, the apparent gain ranging from zero to 30 per cent. 
According to Peabody, a saving of from 5 to 10 per cent may be made by 
jacketing simple and compound condensing engines, and a saving of 
from 10 to 15 per cent by jacketing triple expansion engines of 300 horse- 
power and under. On large engines of 1000 horsepower or more the 
gain, if any, is very small. (Peabody, ''Thermodynamics," p. 400.) 

Other things being equal, the smaller the cylinder and the lower the 
piston speed the greater is the value of the jacket. Experiments show 
no advantage in increasing the jacket pressure more than a few pounds 
above that of the initial steam in the cylinder, and it is usual to reduce 



RECIPROCATING STEAM ENGINES 391 

the pressure in the jackets of the second and succeeding cylinders of 
multi-expansion engines. (Ripper, ''Steam Engine," p. 170.) 

To be effective, jackets should be well drained, kept full of Uve steam, 
and the water of condensation returned directly to the boiler. 

Pumping engines and other slow-speed engines running at practi- 
cally constant load are generally jacketed, but in street-railway work 
and in the majority of manufacturing plants carrying fluctuating load, 
jackets are not considered advantageous. 

Whatever may be the actual economy due to jacketing, there is no 
question but that the jacket greatly influences the action of the steam 
in the cyhnders, and whether beneficially or not depends upon the 
design and construction of the engine. Unless otherwise specified, 
manufacturers usually build their engines without jackets. 

A revival of the steam -jacket for small single cylinder engines is 
quite probable if the exceptional results obtained by Prof. E. H. Miller 
of the Massachusetts Institute of Technology on a Prosser-Fitchburg 
engine are maintained in practice. The engine is simple, non-con- 
densing. The heads and barrel of the cylinders are jacketed with 
steam at throttle pressure. The cylinder has poppet valves, steam and 
exhaust, and is equipped with a double eccentric valve gear. A total 
steam consumption (steam dry and saturated at admission) of 20.46 
lb. per i.hp-hr. was recorded, corresponding to a Rankine cycle ratio of 
83 per cent. With steam superheated to 86.7 deg. fahr. the water rate 
was reduced to 16.59 lb. per i.hp-hr, corresponding to a Rankine cycle 
efficiency of 88.7 per cent. Rankine cycle ratios as high as 92.3 per 
cent are purported to have been reafized in shop tests. See Table 
78 for results of Prof. Miller's Tests. 

Jacketing Applied to Steam Cylinders: Power, Mar. 18, 1913, p. 368. 

184. Receiver Reheaters: Intermediate Relieating. — The receivers be- 
tween the cyhnders of multi-expansion engines are frequently equipped 
with heating coils, as illustrated in Fig. 453, the function of which is to 
superheat the exhaust steam before delivering it to the cylinder im- 
mediately following, with a view of reducing the losses occasioned by 
cylinder condensation. The coils are suppUed with live steam under 
boiler pressure and may serve to evaporate a portion of the moisture 
or to actually superheat the steam supplied to the following cylinder. 
The question of the propriety of using reheaters is an open one, since 
reliable data relative to their use are meager and discordant. The con- 
ditions under which the few recorded tests were made are too diverse 
to warrant definite conclusions. Some show an appreciable gain in 
economy, others a decided loss. A reheater is of little value in improving 



392 STEAM POWER PLANT ENGINEERING 

the thermodynamic action of the engine, and is probably a loss unless 
it produces a superheat of at least 30 deg. fahr., and to be fully ef- 
fective should superheat above 100 deg. fahr. (L. S. Marks, Trans. 
A.S.M.E., 25-500.) The effectiveness of the reheater will evidently be 
increased by the removal of the greater portion of the moisture from the 
exhaust steam before it enters the receiver. In the 5500-horsepower 
engine at the Waterside Station in New York it was shown that both 
jackets and reheaters, either together or alone, were practically value- 
less, throughout the working range of load. (Power, July, 1904, p. 
424.) Many similar cases may be cited which show no gain in economy 
with the use of the reheaters. In all cases the reheater effects a great 
reduction in the condensation in the low-pressure cyhnders, but the 
resulting gain, considering the condensation in the reheater coils, may 
be little, if any. On the other hand, with properly proportioned re- 
heaters, the gain may be considerable and particularly with super- 
heated steam. Practically all European engines operating with highly 
superheated steam are equipped with receiver-reheaters. In the loco- 
mobile type of engine plant the intermediate reheating is effected by 
heating coils placed in the path of the furnace gases. See also para- 
graph 194. 

In triple-expansion pumping engines receiver-reheaters are found to 
effect an appreciable gain in economy, and practically all such engines 
are equipped with them. In electric traction work or where the load 
is a widely fluctuating one the reheater has been virtually abandoned. 
Apart from the consideration of fuel economy, all tests show a marked 
increase in the indicated power of the low-pressure cylinder (5 to 15 
per cent), and to that extent it increases the capacity of the entire 
engine. (G. H. Barrus, Power, Sept., 1903, p. 516.) 

Engine Reheaters: Mech. Engr., Dec. 23, 1910. 

185. Compounding. — If the entire expansion instead of being effected 
in a single cylinder is allowed to take place in two or more cylinders 
the engine is said to be '^ compounded. " The term '^ compound" 
without quahfication, however, refers only to the two-cy Under arrange- 
ment. If expansion takes place in three stages the engine is known 
as a triple-expansion engine; similarly, the four-stage machine is 
called a quadruple expansion engine. When high-pressure steam is 
admitted into a single cylinder engine of the ordinary double-flow type 
and expansion is carried down to a comparatively low point a large 
portion is condensed by the metal surfaces; at the end of the stroke and 
during exhaust some of the water is re-evaporated, but the steam so 
formed is discharged without doing useful work. If the same weight 



RECIPROCATING STEAM ENGINES 393 

of steam is expanded through the same pressure range in a ('()m])()und 
engine, the temperature range in each cyUnder will Ix^ less, initial con- 
densation will be reduced and part of the heat lost in the first cylinder 
by leakage and clearance will do work in the second cylinder. The 
higher the temperature range the more pronounced will be the thermal 
economy effected by compounding. The number of stages is limited 
commercially because of the first cost, complexity, cost of lubrication, 
attendance and maintenance. 

Cylinder ratios for high-speed single-valve compound engines vary 
from about 1 to 2i with 100 pounds pressure to about 1 to 3 with a 
pressure of 150 pounds, and for slow-speed condensing engines from 1 
to 3 with 125 pounds pressure to about 1 to 4 with a pressure of 175 
pounds. G. I. Rockwood recommends a ratio as high as 1 to 7, and a 
number of engines designed along this line have shown exceptional 
economy. For variable load operation two stages appear to give the 
best ultimate economy. In case of very large condensing engines the 
last stage consists of two cylinders because of the unwieldy and costly 
size of a single unit. For constant loads as in pumping stations and 
large marine installations three and four stages appear to be the best 
investment. The ratio of expansion for a multi-expansion engine is 
the ratio of the volume at release in the low-pressure to that at cut-off 
in the high-pressure cylinder. Commercially it is usually taken to be 
the product of the ratio of the volume of large to small cylinder divided 
by the fraction of the stroke at cut-off in the high-pressure cylinder. 
For example, a compound engine with cylinders 24-in., 48-in. l)y 48-in. 
cutting off at J in the high-pressure cylinder has a nominal ratio of ex- 
pansion of 4 -^ § = 12. The number of expansion at rated load in 
multi-expansion condensing engines varies widely, ranging from 10 to 
33, with an average not far from 16. 

The respective advantages and disadvantages of compounding may 
be tabulated as follows: 

Advantages Disadvantagp:s 

1. Permits high range of expansion. 1. Increased first cost due to 

2. Decreased cylinder condensation. multiplication of parts. 

3. Decreased clearance and leakage 2. Increased hulk. 

losses. 3. Increased complexity. 

4. Equalized crank effort. 4. Increased wear and tear. 

5. Increased economy in steam con- 5. Increased radiation loss. 

sumption. 

186. Uniflow or Unaflow Engine. — A study of the Rankine cycle will 
show that the greater the pressure range between admission and release 
the greater will be the theoretical thermal efficiency. In the standard 



394 



STEAM POWER PLANT ENGINEERING 



double-flow, un jacketed type of engine the actual thermal efficiency 
increases with the pressure range up to a certain maximum beyond 




Fig. 208. Section through Cylinder of a Nordberg Uniflow Engine Showing 
Location of Cataract ReUef Valve. 

which increased leakage and cylinder condensation offset the theoreti- 
cal gain. This maximum varies with the type and size of engine, num- 
ber of cylinders, design of valve gear and other influencing factors. 




^^^^^^^M 



^^^^^^^^^^^ 



K\\\\\w^ 




Fig. 209. Section through Cyhnder of a C. & G. Cooper Co.'s. Uniflow Engine. 

Cylinder condensation is increased with the pressure range because 
the cylinder head and other clearance surfaces are chilled by the ex- 
haust steam so that at the beginning of the stroke a considerable portion 



RECIPROCATING STEAM ENGINES 



395 



of the incoming steam is condensed. With superheated steam an 
equivalent heat exchange takes place. 

In the uniflow engine the steam enters at the end of the cylinder as 
in the double-flow type but it is exhausted from the center at the fur- 




FiG. 210. Section throufj;h Cylinder of a Skinner "Universal " Uniflow Engine. 

thest point from the heads. (See Fig. 208.) Consequently the cyhnder 
heads are exposed to exhaust temperature only for the very small length 
of time that it takes the piston to uncover the ports. Furthermore, 
the heads are jacketed with live steam (and in some designs the entire 



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Fig. 211. Performance of 19-inch by 30-inch C. & G. Cooper Co.'s Uniflow 

Engine. 



396 



STEAM POWER PLANT ENGINEERING 



cylinder) so that on the return stroke the steam at exhaust temper- 
ature and pressure is compressed against the hot surfaces of the cylinder 
head; thus when the admission valve opens the incoming steam meets 
no cold surface and cylinder condensation is largely prevented. The 
result is that a wide pressure and temperature range can be allowed 
in a single cyhnder with good economy. In fact the heat consumption 

of a uniflow engine operat- 
ing condensing is equal to 
that of a compound Corliss. 
For ordinary non-con- 
densing service the econ- 
omy is no better than 
that of a high-grade single- 
cyHnder poppet-valve en- 
gine for the same operat- 
ing conditions. This is due 
primarily to the harmful 
effects of the excessive 
pressure resulting from 
compression or to the 
methods employed for re- 
ducing this pressure. In 
the typical uniflow enghie 
of low clearance volume, 
compression begins as soon 
as the piston covers the 
exhaust ports (correspond- 
ing to approximately 90 
per cent of the stroke), and 
for moderate or low vacua 
the resulting pressure at 
the end of compression is equal to or less than that at admission, but 
with high back pressure as in non-condensing service it may be greatly in 
excess. To prevent this excessive rise in compression for non-condens- 
ing service American manufacturers either increase the clearance volume 
(Ames Stumpf ^^Unaflow" engine and C & G. Cooper Co. '^ Uniflow" 
engine) or employ an auxiliary valve which delays compression (Uni- 
versal ^^Unaflow" engine). For high initial pressures the clearance 
volume may be reduced with a resulting increase in economy. In case 
the vacuum is lost when operating condensing, compression pressure 
may be reheved by an adjustable snifting valve (Fig. 208) or by means 
of the auxiliary valve mentioned above. Some idea of the economy 



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Uniflow,Engine 2l"x 22," 200 R.p.m., 140 lb. per.aq. in. 

(Saturated) Steam Pressure 

Generator 325 K.Y.A.. 80-Per Cent P.F., 250 Kw.. 200 


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Fig. 212. Guarantee 'and Test Performance of a 
21-inch by 22-inch Uniflow Engine. 



RECIPROCATING STEAM ENGINES 



397 



TABLE 77. 

OPERATING PERFORMANCE OF A 33-IN. BY 36-IN. C. & G. COOPER CO.'S UNIFLOW 

ENGINE. 

1. Engine, C. & G. Cooper Co., rated output 640 i.hp. or 449 kw. at switchboard. 

2. Cylinder 33 in. by 36 in. Piston rod, head end, 5] in. diameter; crank end, oj in. diameter. 

3. Clearance, head end 4 per cent; crank end 3.8 per cent. 

4. Boilers, Heine Boiler Co., rated at 200 hp. each. 2000 sq. ft. heating surface. Dean shaking grates 

hand fired. Grate area 39.6 sq. ft. 
Auxiliaries, Dean Bros. Surface condenser and pumps. 1400 sq. ft. cooling surface in condenser. 

5. Date, April 25 and 26, 1914. 



6. No. of run . . 

7. Duration min. 

8. Barometric pressure: 

(a) in. of mercury 

(b) lbs. per sq. in 

9. Steam pressure in sup- 
ply pipe lb. 

10. Vacuum referred to 30 
in. barometer: 

(a) at engine in. 

(b) in condenser in. 

11. Temperature of steam 
at engine deg. fahr. 

12. Superheat at engine, 

deg. fahr. 

13. Temperature of con- 
densed steam in meas- 
uring tanks. . . deg. fahr. 

14. Temperature of cooling 
water entering con- 
denser deg. fahr. 

15. Temperature of cooling 
water leaving condenser, 

deg. fahr. 

16. Temperature of engine 
room deg. fahr. 

17. Steam used by engine 
during run lb. 

18. Steam used by engine 
per hr lb. per hr. 

19. R.p.m. of engine 

20. Piston speed of engine, 

ft. per min. 

Power as Measured at 
Switchboard. 

21. Volts 

22. Amperes 

23. Kilowatts by watt- 
meter kw. 

24. Steam u.sed by engine 
under actual conditions 
of operation, lb. per kw-hr. 

Heat Data. 

25. Heat units in each lb. 
of steam supplied . B.t.u. 

26. Total units supplied per 
hr. per kw., 

B.t.u. per kw-hr. 

27. Thermal efficiency ra- 
tio between heat equiv- 
alent of kw. at the 
switchboard and heat 
units supplied in the 
steam per kw 

28. Heat units which 
would be obtained by 
perfect (adiabatic) ex- 
pansion from initial to 
final pressure per lb., 

B.t.u. 

29. Heat units per kw., 

B.t.u. per kw. 

30. Rankine cycle ratio, 

per cent 



1 

191 


2 
62 


29.34 
14.39 


29.35 
14.40 


161 60 


162 50 


23.48 
24.61 


22.67 
24.34 


408.5 


422.1 


37.2 


50.4 


106 


110 


76.9 


77.2 


88.6 


88.4 


80 


78 


26201 


10041 


8230 
124.96 


9722 
124.43 


749.76 


746.58 


242.5 
1048 


243.7 
1212 


440 8 


513.3 


18,67 


18,94 


1108 


nil 


20680 


21060 


16.5 


16.2 


227 


272 


5170 


5150 


66.2 


66.3 



3 
112 

29.36 
14.40 

163.90 

22.55 
24.18 

415.0 

42.8 

113 

73.4 

87.4 

80 

20156 

10798 
123.92 

743,52 



29.40 
14.43 



161.20 



20 98 
23.77 



423,1 
52.3 



72.2 

91.5 
80 



13594 
123.47 



740,82 



243.7 244.5 
1354 1583 



555.1 
19.45 

1106 



569,2 
20.31 



1103 



21510 122400 



15.9 



271 
5270 



15.3 



258 
5250 
65.1 



5 
103 


6 
53 


7 
60 


8 
52 


29.48 
14.48 


29.48 
14.48 


29.50 
14.49 


29.50 
14.49 


158.50 


154.20 


145.60 


125.30 


23 56 
24.18 


19,15 
23,27 


23,26 
23,32 


"24;35 


410.2 


423,1 


412,1 


381.3 


40.3 


55.3 


48.5 


28.3 


96 


126 


92 


83 


74.5 


70.2 


77.2 




82.9 


90.7 


79.6 





79 


81 


86 




10070 


16068 


4070 


2002 


5866 
125.52 


18190 
124.23 


4070 
124.50 


2280 
124.04 


753.12 


747.00 


745,38 


744.24 


243.7 
794 


244.7 
1959 


244.7 
535 


245.7 
270 


336.3 


815.5 


239,2. 


123.2 


17.44 


22.31 


17.02 


18.57 


1110 


1093 


nil 


1098 


19360 


24400 


18900 


20380 


17.6 


14 


18 


16.8 


275 


243 


267 


267 


4800 


5430 


4550 


4960 


71.2 


63.0 


75.0 


68.9 



9 
30 

29.50 
14.49 

127.50 

23.53 
373.3 
19.1 



362 



725 
124.20 



745.20 



398 STEAM POWER PLANT ENGINEERING 

effected by the American types of uniflow engine is shown in Fig. 212. 
In Europe the uniflow engine has been developed to a very high point of 
efficiency and exceptional heat economies have been recorded. Aside 
from the high efficiency in a single cylinder a characteristic feature of 
the uniflow engine is the capacity for heavy overloads and low under- 
loads with a flat water rate curve (Fig. 211). 

The cylinder diameter of the uniflow engine is larger than that of 
an equivalent single cylinder non-condensing engine, but it is smaller 
than the low-pressure cylinder of an equivalent compound engine. 

It is difficult to predict the extent to which the uniflow engine will 
replace the double-flow type, but if the claims of the builders are sub- 
stantiated it will prove a formidable competitor of both the compound 
piston engine and turbine at least for sizes ranging between 200 and 
2000 horsepower. 

187. Use of Binary Vapors. — A consideration of the Carnot or 
Rankine cycles shows that theoretically the eflftciency of the steam 
engine may be increased by raising the temperature of the steam sup- 
plied or by lowering the temperature of the exhaust, that is to say, 
by increasing the range. Superheated steam development has prac- 
tically determined the upper limit, and economical practice indicates a 
vacuum of about 26 inches, corresponding to 126 deg. fahr., as the 
average lower limit for most efficient results from a commercial stand- 
point. 

In the binary-vapor engine the working range has been considerably 
increased by substituting a highly volatile liquid, as sulphur dioxide, 
for the water which is ordinarily used as the coofing medium in the 
surface condenser. 

The SO2 in condensing the exhaust steam is itself vaporized and the 
vapor, under a pressure of about 175 pounds per square inch, used 
expansively in a secondary reciprocating engine. The exhausted SO2 
is discharged into a surface condenser in which it is liquefied by coofing 
water much the same as in refrigerating practice and used over and 
over again. Referring to Fig. 213, which illustrates diagrammatically 
a binary- vapor engine at the Royal Technical High School, Berlin: 
A, B, and C are the three steam cylinders of an ordinary triple-expan- 
sion engine and D the SO2 cyfinder. All four cyUnders drive a common 
crank shaft E. F is sl high-pressure surface condenser which acts as 
a vaporizer for the SO2 and a condenser for the steam. G is a surface 
condenser which serves to condense the SO2 vapor. H is a. liquid SO2 
tank. The operation is as follows: Highly superheated steam enters 
the high-pressure steam cylinder at I and leaves the low-pressure 
cylinder at J, just as in any steam engine. The exhaust steam enters 



RECIPROCATING STEAM ENGINES 



399 



chamber F and is condensed by the hquid SO2 passing through the 
coils. The condensed steam and entrained air are removed from the 
chamber by a suitable air pump. The steam in condensing gives up 
its latent heat to the liquid SO3 and causes it to vaporize. The SO2 







1 

1 


-To Air Pump 




1 — 
1 




SO, Vaporizer and Steam 
- ^^ Condenser 

187'' Absolute H8"f 


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n 


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590 
Stej 








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solute 1 

143.5 R.P.M. 




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1 


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Water 


hIp. 

Cylinder 


1 


Inlet 
62.7° 


im Inlet 


i 


Circulating 
Water 
Outlet 




*" 


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Fig. 213. Diagram of Binary-vapor Engine. 

vapor passes from the coils in chamber F to the SO2 engine D and 
performs work. The exhausted SO2 vapor flows from cyUnder D to 
chamber G, and is condensed by cooling water flowing through a series 
of tubes. The liquid SO2 is collected in liquid tank H and thence is 
pumped into the coils in vaporizer F. The approximate temperatures 
and pressures at different points of the cycle are indicated on the dia- 
gram. 

A number of experiments made by Professor E. Josse in the labora- 
tory of the Royal Technical High School of Berlin on an experimental 
plant of about 200 horsepower gave some remarkable results. A few 
of the tests made with highly superheated steam gave the following 
average figures : 

I.hp. (steam end) 146 . 4 

Steam consumption per i.hp-hour 12.8 

I.hp. (SO2 end) 52.7 

Percentage of power of SO2 engine 35 . 9 

Steam consumption per i.hp-hour of combined engine 9 . 43 

When operating under the most satisfactory conditions a perform- 
ance of 8.36 pounds of steam per i.hp-hour was recorded, correspond- 
ing to a heat consumption of 158.3 B.t.u. per minute. While this is an 
exceptional performance better results have been obtained with the 
uniflow engine and high-grade poppet-valve engine of the double-flow 
type. The binary-vapor engine has not proved to be a commercial 
success because of the high first cost and high maintenance charge. 



400 



STEAM POWER PLANT ENGINEERING 



SO2 does not attack the metal surface of the engine unless combined 
with water, in which case sulphurous acid is formed. There is, how- 
ever, no danger from this cause, since the SO2 being under greater 
pressure effectually prevents leakage of water into the SO2 system. 

The SO2 cylinder requires no other lubrication than the SO2 itself, 
which is of a greasy nature. 

Properties of SOo: Trans. A.S.M.E., 25-181. Binary-vapor Engines: Jour. Frank. 
Inst., June, 1903; Elec. World and Engr., Aug. 10, 1901; U. S. Cons. Reports, 
No. 1139, Sept. 14, 1901; Engr. U. S., Aug. 1, 1903; Sib. Jour, of Eng., March, 1902. 

188. Types of Piston Engines. — A general classification of the vari-' 
ous types of engines used in steam power plant operation is unsatis- 
factory because of the overlapping of the various groups and the fol- 
lowing modifications of a chart devised by Hirshfeld and Ulbricht 
C' Steam Power," p. 92) is merely offered as a general summary of the 
different nomenclatures used in connection with this class of prime 
mover. 



Rotative speed basis 



Longitudinal axis basis 



( High speed 
< Medium speed 
( Low speed 

Vertical 
Inclined 
Horizontal 



Ratio of stroke to 
diameter basis 



Cylinder 

Arrangement 



Slide valve 



TD-slide valve 
J Balanced slide valve 
\T^^ rr. „^o^ u^oJo J 1 Multiported slide valve 

Valve gear basis { I Piston valve 

Corliss valve ( Drop cut-off 
Poppet valve ( Positively operated 
[Single expansion or single engine 

Steam expansion basis J Multi-expansion engine \ T?^'''''^ 
I ( Quadruple 

[Double flow [Standard 



( Short stroke 
) Long stroke 

Single cylinder 

Tandem compound 

Cross compound 

Duplex 

Angle compound 



Steam flow basis 



Operating basis 



I Uniflow 



Crank mechanism J Back acting 



basis 



Initial pressure 
Back pressure 



C High pressure 
< Medium pressure 
( Low pressure 
{ Condensing 
( Non-condensing 



j Trunk 
I Oscillating 



No attempt will be made to describe the various types as outlined 
in this chart further than that incident to the discussion of their relative 
merits for power plant service. 

189. High-speed Single-valve Simple Engines. — This style of engine 
is made in sizes varying from 10 to 500 horsepower. The cyhnder 
dknensions vary from 4-in. by 5-in. to 24-in. by 24-in. and the rotative 
speed from 400 to 175 r.p.m. 



RECIPROCATING STEAM ENGINES 401 

When ground is limited or costly and exhaust steam is necessary 
for heating or manufacturing purposes, the high-speed non-condensing 
engine is most suitable for horsepowers of 200 or less, being compact, 
simple in construction and operation, and low in first cost. For sizes 
larger than this the compound or uniflow engine may prove a better 
investment, except in cases where fuel is very cheap or large quantities 
of exhaust steam are to be used for manufacturing purposes. 

Small high-speed engines are seldom operated condensing, since the 
gain due to reduction of back pressure is more than offset by the extra 
cost of the condenser and appurtenances. 

Engines are ordinarily rated at about 75 per cent of their maximum 
output. For example, a 12-in. by 12-in. non-condensing engine running 
at 300 r.p.m., with initial steam pressure of 80 pounds gauge, is normally 
rated at 70 horsepower, though it is capable of developing 90 horse- 
power at the same speed. 

The steam consumption of high-speed single-valve non-condensing 
engines at full load ranges from 26 to 50 pounds per indicated horse- 
power-hour, depending upon the size of the unit and the conditions of 
operation. An average for good practice is not far from 30 pounds. 
With superheated steam a steam consumption as low as 18 pounds 
per horsepower-hour has been recorded. 

Table 81 gives the steam consumption of a number of single- valve 
high-speed engines running condensing and non-condensing, and Fig. 
215 shows some of the results for different loads. The steam consump- 
tion is fairly constant from 50 per cent of the rated load to 25 per cent 
overload, but for earher loads the economy drops off rapidly. The 
desirability of operating the engine near its rated load is at once ap- 
parent. The curves show a marked economy in favor of the larger 
cylinders, but the engines are not of the same make, and the conditions 
of operation are somewhat different. 

The most economical cut-off for a simple engine is about one-third 
to one-fourth stroke when running non-condensing, and about one- 
sixth when running condensing. 

The performances given in Table 81 are exceptional. It is not ad- 
visable to count on a better steam consumption for this type of engine 
than 30 to 35 pounds of steam per i.hp-hr. 

The curves in Fig. 214 give the performance of a modern, high-grade, 
unjacketed, 15-in. by 14-in., high-speed, single- valve, simple, non- 
condensing engine at various ratings. It is not likely that this type 
and size of engine can be designed to better the results shown in the 
curves for the given conditions. 

In general, when the requirements for exhaust steam are in excess of the 



402 



STEAM POWER PLANT ENGINEERING 



100 






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PERFORMANCE CURVES 

OF 

SIMPLE HIGH-SPEED SINGLE-VALVE ENGINE 

STANDARD TYPE 




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10 20 



75 100 

Indicated Horse Power '^ 

40 50 60 70 80 90 100 110 120 130 

Per cent of Rated Indicated Load 



Fig. 214. Characteristic Performance Curves of a High-grade, Single-cylinder, 
Single-valve, Non-condensing Engine. 



47 
45 
43 

|« 

W37 




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125 



Fig. 215. Typical Economy Curves of High-speed, Single-valve, Non-condensing 
Engines. Saturated Steam. 



RECIPROCATING STEAM ENGINES 403 

steam consumption of a simple non-condensing engine a high-grade eco- 
nomical engine is without purpose. 

190. High-speed Multi-valve Simple Engines. — The steam distribu- 
tion in a single-valve engine may give good economy for a very small 
range in load but may be far from satisfactory for a wide range. This 
must necessarily be so since admission, cut-off, release, and compression 
are all functions of one valve, and any change in one results in a change 
of the others. To obviate the limitations of the single valve, many 
builders design engines with two or more valves. With a two-valve 
engine cut-off is independent of the other events, and with four valves 
all events are independently adjustable. In addition to the flexibility 
of the valve gear, the chief feature of the four-valve engines lies in the 
reduction of clearance volume which is made possible by placing the 
valves directly over the ports. The valves may be of the common 
slide-valve, or rotary type. As a class, four-valve engines are more 
economical than those having a less number of valves. The advantages 
and disadvantages of the four-valve over the single-valve engines may 
be tabulated as below. 





Advantages. 




Disadvantages. 


1. 

2. 
3. 
4. 
5. 


Better steam distribution. 
Better regulation. 
Reduced clearance volume. 
Less valve leakage. 
Better economy. 


1. 
2. 
3. 


Increased number of parts. 
Increased first cost. 
Requires greater attention, 



The steam consumption of a high-speed Corliss non-condensing 
engine at full load varies from 21 to 27 pounds of saturated steam per 
i.hp-hr. (pressure 125-140 lb. gauge) with an average not far from 25 
pounds. With superheated steam the water rate may run as low as 
17 lb. per i.hp-hr. The poppet-valve type appears to be more eco- 
nomical in steam consumption than the Corliss, and a water rate for 
saturated steam as low as 18.9 lb. per i.hp-hr. has been recorded. A 
very high degree of superheat can be used with the poppet-valve type 
and water rates as low as 16 lb. per i.hp-hr. (initial pressure 150 lb. 
gauge, superheat 250 deg. fahr.) are not unusual. The high-speed, four- 
valve engine is usually operated non-condensing. Rankine cycle efficien- 
cies over 80 per cent have been realized with both saturated and super- 
heated steam. An exceptional record for a condensing unit is reported 
by Lentz. With steam at 461 lb. abs. initial pressure and steam tem- 
perature of 1018 deg. fahr. a 100-hp. Lentz unjacketed simple engine 
developed an indicated horsepower on a steam consumption of 5.67 lb. 
per hour. 



404 



STEAM POWER PLANT ENGINEERING 



Fig. 216 gives a comparison between a single-valve and a four-valve 
(Corliss typej high-speed engine, using saturated steam, and though the 
engines differ slightly in size, the conditions of operation were com- 
parable and the marked gain in economy of the latter over the former 
is apparent. Both performances are exceptional, and a 10 to 15 per cent 
greater steam consumption may be expected in average good practice. 

As a general rule single-valve simple engines do not exceed 500 horse- 
power in size for stationary work, whereas 1000 horsepower is not an 
uncomiiion size for the multi-valve type. 



Comparative Economy 
of a 

( A) Single Valve High Speed 

and a 

( B) Four Valve High Speed 
Ncn Condensing Engine 

15 X 14 Reeves Simple ( A) 
16 X 16 Fleming Simple ( B) 




40 50 60 70 80 90 
Per Cent of Rated Load 

Fig. 216. 



100 no 120 130 140 



191. Medium and Low-speed Multi-valve Simple Engines. — A com- 
parison of tests of higl:.- and low-speed single-valve engines irrespective 
of design and construction shows the former as a class to be less eco- 
nomical than the latter. With four- valve engines there is no such dis- 
parity, and the high-speed type has shown just as good economy as the 
slow-speed class. 

Of the various types of simple, low- or medium-speed, four- valve en- 
gines the poppet-valve appears to be the more economical in heat con- 
sumption, but so much depends upon the grade of workmanship that 
general comparisons are apt to lead to error. A comparison of the steam 
consumption of a high-speed, four-valve Corliss and a four- valve poppet 
engine, non-condensing, is shown in Fig. 207. The size and initial 



RECIPROCATING STEAM ENGINES 



405 



pressure are somewhat in favor of the poppet-valve mechanism so 
that the result^ are not strictly comparable but the exceptional economy 
of both types is apparent from the curves. 

The following table taken from the report of Prof. Edw. F. Miller of 
the Massachusetts Institute of Technology gives the results of a 
Fitchburg Prosser single-cyHnder, four- valve, jacketed, non-condensing 
engine which establishes a record for a small simple machine of the 
double-flow type using saturated steam. 



TABLE 78. 

ECO ROM Y TESTS OF A 15-IN. BY 24-IN. FITCHBURG-PROSSER ENGINE. 

Non-condensing. 

Test No. 



Barometer, inches 

Boiler pressure gauge, lb. . . . 

Degrees superheat, fahr 

R.p.m 

Indicated horsepower. 

Steamt per i.hp-hr 

B.t.u. per i.hp. per minute. . . 
Rankine cycle ratio, per cent 



1 


2 


29.5 


29.32 


124.5 


121.6 


86.7 


0.999* 


81.05 


82.04 


54.84 


48.74 


16.59 


19.07 


291.0 


317.0 


81.05 


82.04 



29.32 

101.5 
0.999^ 
80.09 
54.39 
20.46 

341.0 
80.09 



* Quality. 



t Includes jacket condensation. 



The low-speed multi- valve single-cylinder unit ranges in size from 
50 to 3000 horsepower with cylinders varying from 12-in. by 30-in. 
to 48-in. by 72-in. The smaller sizes with trip gear operate at 90 to 
120 r.p.m. and the larger at 50 to 100 r.p.m. Without trip gear, speeds 
of 150 r.p.m. are not uncommon but at this speed they are usually 
classified as high-speed engines. 

A few exceptional performances of this type of engine for saturated 
steam are given in Table 81. For results with superheated steam see 
Table 83. 

193. Compound Engines. — It should be borne in mind that the prin- 
cipal object of compounding is to permit the advantageous use of high 
pressures and large ratios of expansion and consequently this type of 
engine need not be considered for pressures lower than 125 lb. per sq. 
in. gauge. This does not signify that 125 lb. is the limiting pressure 
for compounding; on the contrary, compound condensing engines with 
initial pressures as low as 90 lb. have shown better heat economy than 
simple engines of the same capacity, but the thermal gain for these low 
pressures is usually more than offset by fixed charges and other practical 
considerations. In general, compounding increases the steam econ- 
omy at rated load from 10 to 25 per cent for non-condensing engines 



406 



STEAM POWER PLANT ENGINEERING 



and from 15 to 40 per cent for condensing engines. Compound engines 
range in size from the 100-hp. tandem, single-valve, automatic, high- 
speed, non-condensing unit to multi-valve, cross-compound condens- 
ing units of 4000 hp. or more. Compound engines have been built, 
and are still operating, up to 10,000 hp. rated capacity but the steam 
turbine has practically superseded the piston engine for sizes larger 
than 2000 hp. High-grade compound engines of the full poppet-valve 
type with superheated steam are more economical in steam consump- 
tion at rated load than steam turbines of the same capacity, but first 
cost, size, maintenance and attendance are decidedly in favor of the 



30 



25 



30 





\ 




































\ 

\ 




































• 


\ 












Belative Economy 

ofa 

Simple and Compound 

Non-Condensing High Speed 

Engine 












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\ 






















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Compound 






! 1 




-X 



20 30 40 50 60 70 80 90 100 120 140 160 180 

Fig, 217. Comparison of a Simple and Compound Slide-valve Engine. 



turbine, at least for sizes over 2000 hp. Low rotative speed and re- 
versibility, however, are points in favor of the engine, but the former 
may be offset by the turbine in connection with suitable reduction 
gearing. 

With saturated steam the water rate of the standard type of single- 
valve compound non-condensing engine ranges from 22 to 27 lb. per 
i.hp-hr. at rated load. Since this type of engine permits of only a 
moderate amount of superheat the water rate with superheated steam 
is seldom less than 20 lb. per i.hp-hr. Condensing under a standard 
vacuum of 26 inches reduces the water rate approximately 20 per cent. 

The four-valve compound non-condensing engine has a full load 
water rate, with saturated steam, ranging from 17 to 22 lb. per i.hp-hr., 
and with superheated steam an economy as low as 12 lb. per i.hp-hr. 
has been recorded. Rankine cycle efficiencies as high as 83 per cent 
have been realized for both saturated and superheated steam. 

So much depends upon the initial pressure, degree of vacuum and 



RECIPROCATING STEAM ENGINES 



407 



initial temperature that general figures for condensing practice are 
without purpose. A few special cases are Usted in Tables 81 and 86. 
With saturated steam the best performances are in the neighborhood of 
75 per cent of the theoretical Rankine cycle efficiency, while with highly 
superheated steam 90 per cent of the Rankine cycle efficiency has been 
realized. 

A number of exceptional performances are illustrated in Figs. 218 
to 221. 





































1 


\ 








22 




























A 21,41 X 30 Compound 
B 20,40 X 42 Compound 








































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300 400 500 600 700 800 900 
Indicated Horse Power 

Fig. 218. 



1000 1100 laOO 1300 1400 



27 



26 



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5 20 

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2000 2500 3000 3500 4000 4500 5000 
Gross K.W. Output 
Fig. 219. Economy Test of the 5500-horsepower Three-cylinder Compound I]ngine 

and Generator. 



408 



STEAM POWER PLANT ENGINEERING 



c 


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_ 







15 


20 25 30 
Eeceiver Pressure, Lb. per Sq. In. 


35 


10 


21.5 


25 25.5 26 

Vacuum, In. of Mercury 


26.5 


27 



Fig. 220. Performance of 5500-horsepower Engine under Variable Receiving Pres- 
sures and at Different Vacua. 



193. Triple and Quadruple Expansion Engines. — Triple and quadruple 

expansion engines are still in use where the load is practically constant, 
as in marine and pumping-station practice, but have been abandoned 
in street-railway work where the load fluctuates widely in favor of the 
steam turbine or the two- or three-cylinder compound. Some idea of 
the economy effected by triple-expansion pumping engines may be 
gained from Table 79. A 1000-hp. Nordberg quadruple expansion en- 
gine driving an air compressor at the power plant of the Champion 
Copper Co. is credited with a water rate of 11.23 lb. of saturated steam 
per i.hp-hr., initial pressure 257 lb. gauge. This engine operates in 
the ''regenerative cycle" (see paragraph 460), and the steam consump- 
tion is equivalent to 169.3 B.t.u. per i.hp. per minute and the actual 
thermal efficiency 25.05 per cent.* 

194. The Locomobile. — Although classified under ''steam engines" 
the term "locomobile" apphes to the complete power plant and not 
to the engine only. In Europe this type of plant has been developed 
to a high degree of efficiency, and with very high superheat steam con- 
sumptions as low as 6.95 lb. per i.hp-hr. have .been recorded, corre- 
sponding to a coal consumption of 0.75 lb. coal per brake hp-hr. The 
American type of locomobile is not designed for superheat above 250 
deg. fahr. and the best economies are in the neighborhood of 1 lb. of 
coal per brake hp-hr. 

* Trans. A.S.M.E., vol. 28, p. 221. 



RECIPROCATING STEAM ENGINES 



409 



TABLE 79. 
ECONOMY OF MODERN VERTICAL TRIPLE-EXPANSION PUMPING ENGINES. 

(Official Trials.) 





Type. 


Location. 


Rated 
Capacity, 
Millions 
of U.S. 
Gallons. 


Initial 

Gauge 

Pressure. 


Duty. 


Dry 


Date of 

Test. 


Per Thou- 
sand Lb. 
of Dry 
Steam. 


Per 

One 

Million 

B.t.u. 


Steam 

I.h%. 
Hour. 


5- 2-09 
3-10-10 


Holly 
Holly 
Holly 

Holly 
Holly 
Allis 

Allis 
Allis 
Allis 


Louisville, Ky 

Frankfort, Pa 

Albany, N. Y 

Brockton, Mass 

Cleveland, Ohio 

Boston, Mass 

St. Louis, Mo 

St. Louis, Mo 

Milwaukee, Wis 


24 
20 
12 

6 

2.5 
30 

20 
15 
12 


155.1 
180.2 
153.0 

150.0 
149.6 
185.5 

140.6 
126.2 
124.6 


n95.0 
184.4 
182.1 

170.0 
164.6 
178.5 

181.3 
179.4 
175.4 


164.5 


*9.64. 


4-29-10 






10-14-09 






12- 5-07 
5- 2-00 

2- 4-06 
2-26-00 
1-15-10 


148.8 
163.9 

158.8 
158.1 
151.0 


11.51 
10.33 

10.66 
10.67 
10.82 







* 


109 degrees F. superheat at throttle. 






Date of 
Test. 


Type. 


R.P.M. 


Water Actually 
Pumped, Mil- 
lions of U.S. 
Gallons 24 Hr. 


Net Head 

Pumped 

Against, Lb. 

per Sq. In. 


Indicated 
Horse 
Power. 


Developed 
Horse 
Power. 


Thermal 

Efficiency 

Per Cent. 

I.h.p. 


5- 2-09 
3-10-10 
4-29-10 


Holly 
Holly 
Holly 


24.0 
20.1 
22.3 


24.111 

21.219 
12.193 


90.0 

95.7 

139.5 


925.7 


879.4 
817.0 
726.0 


22.54 


10-14-09 

12- 5-07 

5- 2-00 


Holly 
Holly 
Allis 


40.1 
62.3 

17.7 


6.316 

2.142 

30.314 


130.6 

180.7 

61.0 


i58!7 
801.5 


334.0 
151.9 

747.8 


19.13 
21.63 


2- 4-06 
2-26-00 
1-15-10 


Allis 
Allis 
Allis 


16.5 
16.4 
20.4 


20.070 
15.121 
12.430 


104.0 
127.0 
121.0 


859.2 
801.6 
673.0 


839.6 
726.3 
618.0 


20.92 
21.00 
20.25 



Fig. 221 shows a longitudinal section through a Buckeye-mobile, 
illustrating a well-known American design of locomobile. The entire 
plant is self-contained and requires very Uttle floor space. The engine, 
of the compound center crank type, is set upon the boiler with cylinders 
projecting into the ''smoke-box" so as to minimize piping and radiation 
losses. Steam is generated in an internally fired tubular boiler at a 
pressure of 225-275 lb. per sq. in. gauge and is superheated to a total 
temperature of 600-700 dcg. fahr. Exhaust steam from the high-pressure 
cylinder is reheated by an auxiliary superheater (adjoining the main 
superheater) before it enters the low-pressure cylinder. The feed 
water is heated by an economizer or reheatcr placed in the })recching. 
The condenser is of the jet type and is provided with a rotary air pump. 



410 



STEAM POWER PLANT ENGINEERING 



Surface condensers are installed where conditions necessitate this type. 
All auxiliaries are driven by the main engine. Buckeye-mobiles are 
made in nine sizes ranging from 75 to 600 horsepower, rated capacity, 



Flue gases Lagging lined with 
surrounding cylinder beat insulating material 
High Pressure Cylinder ^ J Low Pressure Cylinder 




Fumaee Gas' 

Fig. 221. Longitudinal Section through a Buckeye-mobile. 

for belt drive or gearing. For direct-connected electric service the 
sizes range from 50 to 400 kilowatts. 

These small plants give over all economies reached only by large 
central stations. 

Fig. 222 shows the performance of a 150-hp. Buckeye-mobile under 



§.16 

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2000 g 
1800 'f 
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1400 M 



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25 50 ?5 100 125 150 175 200 225 

Indicated Horse Power Fuel - Pocahontas Run of Mine.UOOO B.T.U. 

Fig. 222. Economy Test of 150-horsepower Buckeye-mobile, 

various load conditions. Initial pressure 220 lb. gauge, vacuum 25 
in. referred to 30 in. barometers. 

The remarkable economy effected in Europe is shown in Table 80. 

195. Rotary Engines. — The rotary engine differs from the recip- 
rocating engine in that the piston, or equivalent, rotates about the 



RECIPROCATING STEAM ENGINES 



411 



TABLE 80. 

A REMARKABLE ENGINE PERFORMANCE. 

200, 400 X 400 mm. Locomobile. 
(7.8, 15.7 X 15.7 in.) 



Num- 
ber of 
Test. 



Initial Pres- 
sure, Lb. 
per Sq. In. 



Condenser 
Pressure, 
Lb. Abs. 



Steam Temperatures, Deg. Fahr. 



Entering 

High- 
pressure 
Cj'linder. 



Leaving 

High- 
pressure 
Cylinder. 



Entering 

Low- 
pressure 
Cylinder. 



J'inal Feed 
Water. 



R.p.m. 



Condensing with Intermediate Superheating. 



1 


220 


1.47 




Saturated. 




242 


236 


2 


227 


1 . 17 


712 


377 


462 


212 


241 


3 


220 


1 . 17 


718 


367 


460 


206 


242 


4 


221 


1.17 


806 


426 


530 


221 


246 


5 


220 


1 . 17 


842 


469 


538 




243 


6 


220 


1.17 


872 


520 


... 


241 


243 







Non-co 


ndensing without Intermediate Superheating. 




7 


220 




832 




462 


289 


237 


8 


220 




856 




505 


284 


238 


9 


221 




878 




527 


284 


242 


10 


220 




869 




572 


257 


241 


11 


220 




817 




525 


248 


241 


12 


221 





878 




568 


259 


241 



* Compiled from Zeit. des Ver. deut. Ingr., June, 1911. 



No. of 
Test. 



I.Hp. 



D.Hp. 



Mechanical 
Efficiency, 
Per Cent. 



Steam Consumption, 
Pounds. 



Perl.Hp- 
hr. 



Per D.Hp- 
hr. 



Coal Burned, 

Lb. perD.Hp- 

hr. 



Heat Consump- 
tion B.t.u. per 
I.Hp. per 
minute.* 







Condensing with Intermediate Superheating. 




1 


112.5 


103.2 


91.6 


13.98 


14.19 


1.59 


260 


2 


138.4 


132.8 


96.0 


8.51 


8.87 


1.00 


198 


3 


140.3 


131.4 


93 . 5 


8.33 


8.90 


1.00 


195 


4 


140.4 


133 4 


95.0 


7.68 


8.06 


0.96 


186 


5 


138.8 


132.5 


95.5 


7.24 


7 . 56 


0.87 


175 


6 


141.8 


134.0 


94.5 


7.15^ 


7.56 


0.86 


175 



Xon-condensing without Intermediate Superheating. 



7 


61.5 


49.3 


78.0 


11.22 


14.43 


1.65 


262 


8 


83.8 


74.0 


88.0 


10.60 


11.84 


1.17 


249 


9 


111.0 


98.5 


88.0 


9.95 


11.38 


1.12 


237 


10 


129.9 


120.8 


93.0 


10.00 


10.88 


1,07 


238 


11 


140.4 


132.2 


94.0 


10.68 


11.34 


1.12 


248 


12 


142.1 


132.4 


93.0 


9.93 


10.66 


1.05 


235 



Above ideal feed-water temperature corresponding to exhaust pressure. 



412 



STEAM POWER PLANT ENGINEERING 



cylinder axis. Its operation is entirely different from that of the 
steam turbine; in the rotary engine the static pressure of the steam 
actuates the piston and in the turbine the momentum of the steam is 
imparted to the rotating element. 

Over 2200 patents have been issued to date on rotary engines but 
not a single machine has yet been able to compete with the reciprocat- 
ing engine as regards steam economy. The advantages of the rotary 
engine are many and for this reason innumerable inventors have been 
exerting their skill in the development of this type of prime mover, 
but unfortunately the impracticability of satisfactorily packing the 
rubbing surfaces has more than offset the advantages and the com- 
mercially successful machine is yet to be found. 




Fig. 223. Herrick Rotary Engine. 



The writer has tested out various types of rotary steam engines, and 
the best has been but a poor competitor of the ordinary grade of re- 
ciprocating mechanism. 

One of the most successful rotary engines is illustrated in Fig. 223. 
The device consists essentially of two rotors in rolling contact, the 
upper one containing a recess which serves as a steam inlet and allows 
the piston on the lower rotor to pass, while the lower one contains the 
piston and transmits the power to the shaft. In fundamental prin- 
ciple it is not unlike many other rotary engines in that the power is 
applied directly to the shaft by the expansion of steam behind a rotary 
piston. The synchronous movement of the two rotors is maintained 
by means of two timing gears on the far side of the casing. The curves 



RECIPROCATING STEAM ENGINES 



413 



in Fig. 224 are based upon the tests made by Professor Pryor of Stevens 
Institute of a 20-horsepower engine of this design, initial pressure 150 
pounds gauge, atmospheric exhaust, steam dry and saturated. 



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Atmospheric Back Pressure, 

Steam,Dry and Saturated 

1000 R.P.M. 




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neoo 



1400 



1200 



1000 1} 



800 



600 



400 



300 



2 4 6 8 10 12 14 16 18 20 ti 24 26 
Brake Horse Power 

Fig. 224. Performance of Rotary Engine. 

196. Throttling vs. Automatic Cut-Off. — The action of the gov- 
ernor in the throttUng engine is shown by the superposed indicator 
cards (Fig. 225) taken between zero or friction load and maximum 
load. The effect of throttling is to reduce the pressure during admis- 




FiG. 225. Typical Indicator Cards. High-speed Throttling Engine. 

sion, but does not change the point of cut-off or other events of the 
stroke. The steam may be partially dried or even superheated by 
throttling, thus tending to reduce cyUnder condensation. Initially dry 



414 STEAM POWER PLANT ENGINEERING 

saturated steam at a pressure of 125 pounds gauge would be super- 
heated about 12 degrees in expanding through a throttle to 90 pounds, 
or if it contained initially 2 per cent moisture would be perfectly dried 
in expanding to 40 pounds. Friction through the valve also tends to 
dry the steam. Thus with very Hght loads the superheat may be 
appreciable. The possible gain due to decreased cylinder conden- 
sation is to some extent offset by incomplete expansion. The best 
efficiency for a given load is realized by a proper compromise between 
cut-off and initial pressure. Experiments made by Professor Denton 
(Trans. A.S.M.E., 2-150) on a 17-in. by 30-in. non-condensing double- 
valve engine showed the most economical results with J cut-off for 
90 pounds pressure, | cut-off for 60 pounds, and yVo for 30 pounds. 
The average throttling engine does not give close regulation, the governor 
usually lacking sensitiveness. Tests show the economy to be better 
than that of the automatic engine on Hght loads, and the crank effort 
more uniform. 

The indicator cards shown in Fig. 226 were taken from a single- 
valve high-speed automatic engine operating between friction load and 
maximum load. The mean effective pressure is adjusted to suit the 
load by the automatic variation in the cut-off, the initial pressure 




Fig. 226. Typical Indicator Cards. High-speed Automatic Engine. 

remaining the same. Since the cut-off is controlled by the action of 
the governor on the single valve, all other events of the stroke are like- 
wise changed. With a four-valve engine the variation in cut-off does 
not affect the other events. 

The chief advantage of the automatic over the throttling engine lies 
in its sensitive regulation, and while, in general, it gives a lower steam 
consumption than the throttling engine, this is probably in most cases 
due to superior construction and not to the method of governing. 

The following performances of a Belliss 250-horsepower high-speed 
condensing engine fitted with both automatic and throttling governing 
devices give results decidedly in favor of the throttling engine. (Pro. 
Inst, of Mech. Engrs., 1897, p. 331.) 



RECIPROCATING STEAM ENGINES 



415 



Percentage of load. . . 
Electrical horsepower 
Steam per i.hp-hour. 



A 


atomatic Cut-Oflf 






Throttling. 


100 


62.5 


33 


25 


100 


62.5 


33 


213 


132 


77.8 


53 


213 


132 


77.8 


22.0 


22.9 


28.5 


34.3 


21 


21.7 


25.6 



25 
53 

28.4 



Some of the comparative advantages and disadvantages of the auto- 
matic and throttling engines are as follows: 



Automatic. 

1. Sensitiveness of regulation. 

2. Increased ratio of expansion. 

3. Low terminal pressures. 



Throttling. 
Advantages. 

1. Low first cost. 

2. Crank effort more uniform. 

3. Reduced cylinder condensation. 

4. Simplicity of regulating device. 



Disadvantages. 

1. Increased cylinder condensation. 1. Low ratio of expansion. 

2. Greater variation in crank effort. 2. High terminal pressure. 

3. Complicated valve gear. 3. Low initial pressure at early loads. 

4. Low economy at very early loads. 

Fig. 227 shows the relative steam consumption of an engine under 
the same conditions of load when controlled by variable expansion and 
by throtthng. Suppose this en- 
gine to be altered in capacity so 
that the m.e.p. referred to the 
low-pressure piston is about 32, 
then the steam consumption with 
the throtthng governor will be 
as shown by straight hne A. 
This shows that between 32 and 
12 pounds m.e.p. very httle is 
gained by a variable expansion, 
and below that load the throttled 
governor is the more economical. 
(Power, Feb. 21, 1911, p. 301.) 

197. Selection of Type. — Modern operating conditions are so diversi- 
fied and at the same time so speciaUzed that the selection of the type 
best suited for a proposed installation is an increasingly difficult prob- 
lem. That engineers are not agreed as to the best practice is evidenced 
by the different types of engines selected for practically identical oper- 
ating conditions. General rules are without purpose since each par- 
ticular installation is a problem in itself. Floor space, capacity, cost 
of fuel, water rate, steam pressure, water supply, load characteristics, 




Mean Pressure on L.P.PjBton, Lb.per Sq.In. 

Fk;. 227. Throttling !vs. Automatic Cut-off. 



416 STEAM POWER PLANT ENGINEERING 

exhaust steam requirements, size of foundation, vibration, first cost, 
attendance and maintenance all govern the selection of type. The 
principal factor governing the size of units is the station load curve 
or rather, load curves. Where these load curves are known the problem 
is a comparatively simple one, but when they must be assumed as is 
generally the case with a new project, it is largely a matter of experience. 

How to Select Prime Movers for Industrial Electrical Electric Generating Plants: 
Eng. Mag., Aug., 1916, p. 705. 

Economic Selection of Prime Movers: Power, Oct. 12, 1915, p. 511. 

198. Cost of Engines. — The cost of engines like any other commodity, 
varies with the price of raw material, cost of labor, design and grade of 
workmanship. Even a list of current prices is subject to discount in- 
cident to competition. Consequently data of this nature can only be 
of a very general nature and must be used accordingly. Specific 
figures should be obtained from builders if accurate comparisons are to 
be made between the various types for actual installation work. Prices 
per rated horsepower range from $4.00 to $25.00 and per pound from 
5 to 15 cents. The more economical engines are generally higher 
priced. The following rules are based on average costs and should not 
be used except for rough estimates. 

Simple high-speed engine C = 300 + 8 X i.hp. 

Compound high-speed engines C = 1000 + 15 X i.hp. 

Simple low-speed engines C = 1000 + 10 X i.hp. 

Compound low-speed engines C = 2000 + 13 X i.hp. 

Other rules given in this connection by different authorities are as 
follows : 

Simple high-speed engines C = 435 + 6 . 5 X i.hp. 

Simple non-condensing Corliss C = 700 + 10 X i.hp. 

Compound high-speed Corliss C = 500 + 10 . 5 X i.hp. 

Compound Corliss engines C = 1800 + 13 . 6 X i.hp. 

C = cost in dollars, f.o.b. shipping point, 
i.hp. = rated indicated horsepower. 
Cost of setting high-speed engine = 60-4- 0.75 X i.hp. 
Cost of setting low-speed engine = 500 + 1.3 X i.hp. 



RECIPROCATING STEAM ENGINES 



417 



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(418) 



RECIPROCATING STEAM ENGINES 



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(421) 



422 



STEAM POWER PLANT ENGINEERING 



TABLE 86. 

WATER RATES OF PISTON ENGINES AT VARIABLE LOADS. 



SATURATED STEAM. 



Indicated Horse- 


Pounds Steam per Indicated Horsepower Hour at 


power. 


Full Load. 


Three-quarter Load. 


One-half Load. 


One-quarter Load. 



A. Automatic Single-cylinder Non-condensing. Initial Pressure 125- Lb. Gauge: Cut-off 



80 


29.42 


29.93 


31.66 


37.08 


100 


28.96 


29.40 


31.04 


36.00 


125 


28.47 


28.84 


30.42 


35.10 


150 


28.12 


28.46 


29.95 


34.38 


200 


27.51 


27.81 


29.25 


33.20 


300 


26.64 


26.75 


27.97 


31.68 



B. Medium-speed Four- valve Non-condensing. Initial Pressure 130-lb. Gauge: Cut-off \. 



200 


23.45 


23.05 


24.75 


35.00 


350 


23.03 


22.54 


24.07 


33.79 


500 


22.61 


22.06 


23.45 


32.73 


650 


22.24 


21.67 


22.92 


31.71 


800 


22.00 


21.40 


22.57 


30.91 


900 


21.90 


21.31 


22.44 


30.48 



C. Automatic Tandem-compound Non-condensing. Initial Pressure 140-lb. Gauge: Cylinder Ratio 

4 to 1: Cut-off i. 



100 


24.04 


25.06 


29.54 


43.84 


150 


22.94 


23.82 


28.00 


41.40 


200 


22.36 


23.19 


27.19 


40.46 


250 


21.98 


22.75 


26.65 


39.1-3 


350 


21.48 


22.18 


25.96 


38.01 


450 


21.27 


21.92 


25.61 


37.45 





D. Same as C but Condensing. Vacuum 26-in. 




150 


20.25 


21.51 


26.00 


37.83 


300 


19.10 


20.12 


24.10 


33.90 


400 


18.55 


19.45 


23.11 


32.07 


500 


18.15 


18.93 


22.44 


30.90 


600 


17.92 


18.67 


22.08 


30.20 


700 


17.83 


18.55 


21.91 


29.95 



E. Four-valve Medium-speed Compound Non-condensing. Initial Pressure 150-lb. Gauge: Cylinder 

Ratio 4 to 1: Cut-off i. 



300 


19.71 


21.74 


26.82 


39.54 


450 


19.20 


21.10 


25.90 


37.80 


600 


18.91 


20.66 


25.20 


35.46 


750 


18.74 


20.41 


24.73 


35.31 


850 


18.68 


20.32 


24.55 


34.81 


950 


18.66 


20.30 


24.48 


34.47 



F. Same as E but Condensing. Vacuum 26-in. 



300 


15.42 


15.30 


17.26 


24.00 


500 


14.74 


14.60 


16.45 


22.47 


700 


14.29 


14.10 


15.78 


21.30 


900 


13.97 


13.76 


15.34 


20.48 


1100 


13.73 


13.51 


15.02 


19.94 


1300 


13.56 


13.33 


14.83 


19.66 


1500 


13.49 


13.23 


14.71 


19.52 



* Guaranteed performance of a well-known line of high-grade piston engines. 



RECIPROCATING STEAM ENGINES 423 



PROBLEMS 

1. A 40-hp. non-condensing piston engine uses 500 lb. of saturated steam per hour 
when running idle and 1600 lb. per hour when operating at full load; initial pressure 
115 lb. abs. Draw the unit water rate curve assuming that the total water rate 
follows the "Willans" straight-line law. 

2. A 15-inch by 18-inch poppet-valve engine uses 18.8 lb. steam per i.hp-hr. at 
rated load, initial pressure 145 lb. absolute; back pressure lb. gauge; initial quality 
99 per cent; release pressure 4 lb. gauge; mechanical efficiency at rated load 91 per 
cent. Required (on both i.hp. and br.hp. basis): 

a. Heat consumption per hp-hr. 

6. Thermal efficiency, per cent. 

c. Rankine cycle ratio, per cent. 

d. Cylinder efficiency, per cent. 

3. The Rankine cycle ratio of a compound poppet-valve engine is 90 per cent at 
full load; initial pressure, 150 lb. abs.; temperature of steam at admission, 450 deg. 
fahr.; back pressure 16.1 lb. abs. Calculate the full load water rate, lb. per i.hp-hr. 

4. If the exhaust from the engine in Problem 3 is used for heating purposes, re- 
quired the full load water rate, lb. per i.hp-hr. chargeable to power. 

5. A simple engine indicates 160 horsepower on a dry steam consumption of 31 
lb. per i.hp-hr.; initial pressure 130 lb. abs., back pressure lb. gauge. By short- 
ening the cut-off, and by reducing the back pressure to 4-inch mercury (referred to a 
30-inch barometer) the water rate is reduced to 22 lb. per i.hp-hr., the load remaining 
the same. If the condensing equipment requires 10 per cent of the steam supplied 
to the engine for its operation, required the net gain or loss in heat consumption 
per i.hp-hr. due to condensing. 

6. WTiich is the more economical from a heat consumption standpoint, a simple 
non-condensing engine using 26 lb. dry steam per i.hp-hr., initial pressure 100 lb. 
absolute, or a compound condensing engine using 12 lb. steam per i.hp-hr., initial 
pressure 290 lb. abs., superheat 350 deg. fahr., back pressure 2-inches mercury? 
Which is the more perfect of the two? 

See also Problems at end of Chapters XXII-XXIV. 



CHAPTER X 

STEAM TURBINES 

199. Classification. — The development of the steam turbine during 
the past decade has been truly remarkable. So rapid has been the 
growth that many turbines representative of the best practice four 
years ago are virtually obsolete to-day. Because of the almost radical 
changes from year to year it is practically impossible to keep the de- 
scriptive features of a textbook strictly in accord with current practice, 
and the subject matter must necessarily be of a general nature. 

Steam turbines are now being used for driving alternating-current 
generators, turbo-compressors, pumps, blowers and marine propellers, 
and, by means of gearing to furnish power for reciprocating air com- 
pressors, rolling mills and other classes of slow-speed machinery. Al- 
though the reciprocating engine will probably continue to be an impor- 
tant factor in the power world for years to come, its field of usefulness 
is being gradually limited by the steam turbine. The steam turbine 
has found favor chiefly on account of its low first cost, low maintenance 
cost, small floor-space requirements and low cost of attendance. 

A general classification of steam turbines is unsatisfactory because 
of the overlapping of the various groups, and the following chart is 
offered merely as a guide in arranging a few well-known turbines ac- 
cording to the fundamental principles involved in their operation. 



" Impulse 



Steam 

Turbines ] 



Reaction. 

Combined 
Impulse and 
Reaction. 



Single Velocity. { De Laval. 

[Terry. 

I Sturtevant. 



Multi-velocity 
Stage. 



Single-velocity 
Stage. 

Multi-velocity 
Stage. 



Multi-velocity 
Stage. 

Multi- velocity 
Stage. 

424 



I Curtis (Small Type). 
IWestinghouse 

( Kerr. 

< De Laval. 

( Rateau. 



< Curtis. 



rWestinghouse- 
J Parsons. 

I AUis-Chalmers- 
'. Parsons. 

Westinghouse 



Single- 
pressure 
Stage. 



Multi- 

> pressure 

Stage. 



< 



STEAM TURBINES 425 

As shown in the preceding chart, all turbines may be divided into 
three general classes, (1) impulse, (2) reaction, and (3) combined im- 
pulse and reaction, though strictly speaking all turbines depend more 
or less upon both impulse and reaction for their operation. 

Impulse Type: 

In the impulse type the steam is expanded by suitable means and the 
heat given up by the pressure drop imparts velocity to the jet itself. 
The jet impinges against the vanes of a rotating wheel and gives up its 
kinetic energy to the wheel. If the entire pressure drop takes place in 
one set of nozzles and the resulting jet is directed against a single wheel 
the turbine is classified with the single-stage single-velocity group. The 
velocity of the jet is very high, from 2000 to 4000 feet per second, and 
for satisfactory economy the peripheral velocity of the wheel must also 
be very high, from 700 to 1400 feet per second. The De Laval ''Class 
A" turbine is the best-known example of this group. 

If the entire pressure drop takes place in a single set of nozzles and 
a single wheel is to be used at a comparatively low speed satisfactory 
economy may be effected by compounding the velocity. That is, the 
jet issuing from the nozzle at a very high velocity is reflected back and 
forth from the vanes on the rotor to a series of fixed reversing buckets 
until all of the available kinetic energy of the jet has been imparted 
to the wheel. The Terry single-stage turbine is representative of this 
group. 

Low peripheral velocity and high efficiency may be obtained by 
pressure compounding; that is, expansion takes place in a series of 
successive nozzles instead of one nozzle. Only a part of the available 
heat energy is converted into kinetic energy in each set of nozzles. For 
each set of fixed nozzles there is a corresponding rotor. This type of 
turbine is to all intents and purposes a series of single-velocity impulse 
turbines placed side by side. The Kerr turbine is representative of 
this group. 

By compounding both velocity and pressure we have the multi-velocity 
and pressure type of which the Curtis turbine is the best-known example. 

Reaction Type: 

In the reaction type the conversion of potential to kinetic energy 
takes place in the moving blades as well as in the fixed blades. Only a 
very small portion of the heat energy imparts velocity in the first set of 
fixed blades or nozzles. The jet issuing from this set of nozzles impinges 
against the first set of moving blades at a velocity substantially that of 
the moving blades so that it enters them without impulse. The moving 
blades are proportioned so that partial expansion takes place within them 
and the resulting increase in velocity exerts a reaction upon the moving 



426 STEAM POWER PLANT ENGINEERING 

blades. The expansion is very gradual and a large number of alternately 
fixed and revolving blades are necessary to effect complete expansion. 
Because of the small pressure drop in each stage low peripheral veloc- 
ities are possible with high over-all efficiency. The Westinghouse and 
AUis-Chalmers designs of the Parsons turbine are the best-known ex- 
amples of this type. 

Combined Impulse and Reaction Type: 

In this class the high-pressure elements are of the impulse type and 
the low-pressure elements of the reaction type. The Westinghouse- 
Parsons double-flow high-pressure turbine is typical of this class and 
is virtually a combination of the Curtis and Parsons designs. Several 
European impulse turbines as recently designed are fitted with reaction 
blades adjacent to the nozzles, showing the tendency to merge the 
different fundamental types. 

Turbines may be classified according to the service for which they are 
intended, as 

High-pressure non-condensing, 

High-pressure condensing, 

Low-pressure, 

Mixed-pressure, 

Bleeder. 

Each of these types is discussed later on in the chapter. 

Recent Developments in Steam Turbine Practice: Mech. Engr., Jan. 26, 1912. 
The Present State of Development of Large Steam Turbines: Jour. A.S.M.E., May, 
1912. 

The Steam Turbine: Engng., Dec. 29, 1911. 

Status of the Small Steam Turbine: Power, Jan. 2, 1912. 

200. General Elementary Theory. — A given weight of steam at a 
given pressure and temperature occupies a certain known volume and 
contains a known amount of heat energy. If the steam is permitted to 
expand to a lower pressure without receiving additional heat or giving 
up heat to surrounding bodies it is capable of doing a certain amount 
of work which will be the same whether the expansion takes place in 
the cylinder of a reciprocating piston engine, a rotary piston engine, or 
the nozzles and blades of a steam turbine. 

Let W = weight of steam, lb. per sec, 

E = energy given up by 1 lb. of steam, ft-lb., 
Pi = initial pressure, lb. per sq. in. abs., 
Pn = final pressure, lb. per sq. in. abs., 
Hi = initial heat content per lb., B.t.u., 
Hn = final heat content per lb., B.t.u. 



STEAM TURBINES 427 

Then the heat drop, or heat available for doing useful work, is 

W {H, - Hn) B.t.u. (152) 

If the steam expands against a resistance, as, for example, the piston 
of a reciprocating engine, the energy given up in forcing the piston for- 
ward may be expressed 

E, = 777.5 W {H, - Hn) ft-lb. (153) 

If the steam expands within a perfect nozzle the energy will be given 
up in imparting velocity to the steam itself, thus: 

E, = W^{t-]h., (154) 

in which ^ 

Vi »= velocity of the jet in feet per second. 

If the velocity of the jet is retarded to Vn feet per second, as by placing 
a series of vanes in its path, then the energy given up to the vanes 
(neglecting all losses) is 

E = W \g ■ (155) 

If the kinetic energy is completely absorbed by the vanes (neglecting 
all losses), then y„ = and the energy given up is 

E^ = W ^'. (156) 

ButE'i = E%. Hence, 

777.5 W {Hi -Hn) = W^^ 

from which 

Vi = 223.8 VHi - Hn* (157) 

If there are n pressure stages, then the theoretical stage velocity is 



7/ = 223.8 J— — . (158) 

The jet issuing from the nozzle is capable of exerting an impulse 
equal to F upon any object in its path, thus: 

WV^ 

F = -!^1^ lb. (159) 

g 

If A = the area of cross section of the jet in square feet, and y = 
weight of steam, pounds per cubic foot, then W = yAVu or 

F = ^^^"^^ lb. (160) 

9 

* For most purposes it is sufficiently accurate to make 223.8 = 224. 



428 STFAM POWER PLANT ENGINEERING 

The reaction, R, of the jet against the nozzles is equal in value and 
opposite in direction to the impulse, or 

i2 = /r = mi = 1^111-. (161) 

9 9 

The theoretical horsepower developed by a jet of steam flowing at 
the rate of one pound per second may be expressed 



in which 



f] V 2 _ y 2 

HP- = 550 =173^' <'''' 



Vi = initial velocity of the jet, ft. per sec, 
Vn = final velocity of the jet, ft. per sec. 

Steam consumption per horsepower hour: 



Tr.=|5?. (163) 



Heat consumption, B.t.u. per horsepower, per minute: 

Wx (H, - g„) 



60 
in which 




\i.\n) 


Qn = heat of the liquid corresponding to temperature of the exhaust 


Impulse efficiency of the jet = equation (155) - 


7- equation 


(156). 






(165) 


Thermal efficiency (Rankine cycle) : 






^ Hi — Hn 

Rankine cycle ratio : 

p _ 2546 • 




(166) 
(167) 


TFi (ff , - Hn) 



Equations (152) to (167) are general and are applicable to all turbines 
of whatever make. 

The more important types of turbines will be discussed separately 
and an application of above equations will be made in each specific case. 

Heat Drop in Steam Turbines: Trans. A.S.M.E., Vol. 33, p. 325, 1911; Engr., 
Mar. 8, 1912; U.S. Bureau of Standards, Reprint No. 167, 1911. 



I 



STEAM TURBINES 



429 




430 



STEAM POWER PLANT ENGINEERING 



301. The De Laval Turbine. — Fig. 228 shows a section through a 
De Laval steam turbine and gear case and illustrates the principles of 
the single-stage '^ impulse" type. The turbine proper, to the right of 
the figure, consists of a high-carbon steel disk C fitted at the periphery, 
with a single row of drop-forged steel blades and inclosed in a cast-steel 
casing. The disk is secured to a light flexible shaft and is of such a 
cross section that the radial and tangential stresses throughout its mass 
are of constant value. A flexible shaft is employed which allows the 
wheel to assume its proper center of rotation and thus to operate like 
a truly balanced rotating body.* The shaft is supported by three, 
bearings, P, K, and 7. 7 is self -aligning and carries the greater part 
of the weight of the disk. K is a flexible bearing, entirely free to 
oscillate with the shaft, and its only function is to seal the wheel casing 

against leakage. The power 
is transmitted through a steel 
helical pinion K' mounted on 
the extension of the turbine 
shaft E, to two large gears M, 
M at a reduction in speed of 
about 10 to 1. The blades. 
Fig. 229, are made with a bulb 
shank and fitted in slots milled 
in the rim of the wheel. The 
flanges, at the outer end of 
the blades, are brought in 
contact with each other and 
calked so as to form a continuous ring. The inlet and outlet angles of 
the blades are made alike and are 32 degrees for smaller sizes and 36 
degrees for larger sizes. 

The operation is as follows: Steam enters the steam chest D, Figs. 
228 and 230, through the governor (shown in detail in Fig. 231) and is 
distributed to the various adjustable nozzles, varying in number from 
1 to 15 according to the size of turbine. In the earher types the nozzles 
were uniformly distributed around the circumference, but in the later 
types are arranged in groups. As illustrated in Fig. 230 the nozzles are 
placed at an angle of 20 degrees with the plane of the disk. The steam 
is expanded adiabatically in the nozzles to the existing back pressure 
before it impinges at high velocity against the blades. After giving 
up its energy the steam passes into chamber W, Fig. 228, and out 
through the exhaust opening. Fig. 231 gives the details of the governor 

* The shaft diameter for a 100-horsepower turbine is but 1 inch and for a 300- 
horsepower turbine approximately 1 }f inches. 




Fig. 229. De Laval Blades. 



STEAM TURBINES 



431 



and vacuum valves. Two weights B are pivoted on knife edges A 
with hardened pins C bearing on the spring D. E is the governor body, 
fitted in the end of the gear-wheel shaft K, and has seats milled for the 
knife edges A. The spring seat D is held against pins A by spiral 



'^///////////////////////////^^^ 




Fig. 230. De Laval Nozzle. 







jj !■■■ I « tr-iZ.ZZl7 




Fig. 231. De Laval Governor for Single-stage Turbine. 



concentric springs, the tension on which is adjusted by a milled nut /. 
When the speed exceeds the normal, centrifugal force causes the weights 
to fly outward and overcome the resistance of the springs. This pushes 
pin G against bell crank L, which in turn closes the double-seated valve, 
thus throttling the supply of steam. To prevent racing in case the 
load is suddenly removed the vacuum valve T is added to the governor 



432 STEAM POWER PLANT ENGINEERING 

mechanism. Its operation is as follows: The governor pin G actuates 
the plunger H under normal conditions without moving the plunger 
relative to the bell crank. In case the load is suddenly removed, cen- 
trifugal force pushes pin G against bell crank L until it reaches its 
extreme position and the valve is nearly closed and little steam enters 
the turbine. If this does not check the speed, plunger G overcomes 
the resistance of spring M, and H moves relative to L, and its adjustable 
projection presses against valve stem T and allows air to rush into the 
turbine through passage P. 

The power of the turbine depends upon the number of nozzles in 
action, and these can be opened or closed by a hand wheel on each. 
Each nozzle performs its function as perfectly when operating alone as 
when operating in conjunction with others. 

De Laval turbines of the single-stage geared type are made in sizes 
ranging from 17 to 700 horsepower, condensing and non-condensing, 
and are designed to regulate within an extreme variation of 2 per cent 
from no load to full load. The speeds vary from 10,600 r.p.m. for the 
largest size to 30,000 r.p.m. for the smallest, the gearing reducing these 
to 900 and 3000 r.p.m., respectively, at the shaft. The diameter of 
the wheel varies from 4 inches in the smallest turbine to 30 inches in 
the largest, thus giving peripheral velocities of from 520 to 1310 feet 
per second. 

The single-stage geared type just described is no longer manufactured 
by the De Laval Co. and the multi-velocity stage machine is used in 
its place. 

This company also manufactures a multi-pressure impulse turbine. 

Both of these types are described further on. 

202. Elementary Theory. — Single- wheel Single-stage Turbine. — The 
maximum theoretical power developed by a jet of steam flowing through 
a nozzle is dependent only upon the weight of steam flowing per unit of 
time and the initial velocity. Therefore the higher the initial velocity 
for a given rate of flow the greater will be the power developed and the 
higher the efficiency. 

The maximum weight of steam discharged through a nozzle of any 
shape and for a given ini'tial pressure is determined by the area of the 
narrowest cross section or throat. 

To obtain the maximum velocity at the exit or mouth, for a given rate 
of flow, the nozzle should be proportioned so that expansion to the 
external pressure into which the nozzle delivers shall take place within 
the nozzle itself. If expansion in the nozzle is incomplete, sound waves 
will be produced and there will be irregular action and loss of energy. 
On the other hand, if expansion in the nozzle is carried below that of 



STEAM TURBINES 



433 



the external pressure at the mouth, sound waves will be produced with 
subsequent loss of energy even greater than in the former case. 

Experimental and mathematical hivestigations indicate that the pres- 
sure at the narrowest section of an orifice or the throat of a nozzle 
through which steam is flowing falls to approximately 0.58 of the initial 
absolute pressure (with resultant velocity of about 1400 to 1500 feet per 
second) and any further fall in pressure must take place beyond the 
narrowest section. Thus for back pressures greater than 0.58 of the 
initial (conveniently taken as |), maximum exit velocity may be ob- 
tained from orifices of nozzles of uniform cross section or with sides 
convergent. For back pressure less than 0.58 of the initial the nozzle 
must first converge from inlet to throat and then diverge from throat to 
mouth in order to obtain maximum velocity. Without the divergent 
portion of the nozzle the jet will begin to spread after passing the throat, 
and its energy will be given up in directions other than that of the 
original jet. 




Fig. 232. Theoretically Proportioned Expanding Nozzle. 

Fig. 232 shows a section through a theoretically proportioned ex- 
panding nozzle. The cross section of the tube at any point n may be 
calculated by means of equation 

An = ^^ (168) 

in which ^ » 

An = area in square feet, 

W = maximum weight of steam discharged, pounds per second, 
Sn = specific volume of the steam at pressure Pn. 
For wet steam Sn = XnUn + o-, 
in which 

Xn = quality of steam at pressure P„ after adiabatic expansion from 

pressure Pi, 
Un = specific volume of saturated steam at pressure Pn, 
a = volume of 1 lb. of water corresponding to pressure Pn. This 

quantity is very small compared with that of the steam and 

may be neglected. 



434 STEAM POWER PLANT ENGINEERING 

For superheated steam, see Mollier diagram, paragraph 451. 
Vn = velocity of the jet, feet per second. 
Vn may be determined from equation (157) : 



Vn = 223.8 VHi - Hn- 

By substituting Hn = heat content corresponding to pressure Pn 
= 0.58 Pi in equations (157) and (168) the area at the throats may be 
readily determined. The cross-sectional area for other points in the 
tube may be determined in a similar manner by assigning values of 
Hn corresponding to the various pressures. 

In case of a perfect nozzle Hi — Hn represents the heat given up 
toward producing velocity by adiabatic expansion from pressure Pi 
to Pn- In the actual nozzle the frictional resistance of the tube serves to 
increase its dryness fraction, but in doing so it decreases the amount of 
energy the steam is capable of giving up towards increasing its own 
velocity. If y one-hundredths of the heat Hi — Hn is utihzed in over- 
coming frictional resistance, then the resulting velocity will be 

V = 223.8 V(l - y) {Hi - Hn). (169) 

The quality of the steam after expanding to Pn against the resistance 
will be higher by an amount 

In = increase in quality = —i (170) 

in which 

Vn = heat of vaporization at pressure Pn. 

The curves in Fig. 233, calculated by means of equations (157) and 
(168), show the relationship between velocity, quaUty, pressure, and 
kinetic energy for all points in a theoretically perfect nozzle expanding 
one pound of dry steam per second from an initial absolute pressure of 
190 pounds to a condenser pressure of one pound. 

The curves in Fig. 234 are based upon the experiments of Gutermuth 
(Zeit. d. Ver. Ingr., Jan. 16, 1904) and show the effect of a few shapes 
of nozzles and orifices on the actual weight of steam discharged for 
various rates of initial and final pressures, the smallest section of the 
tube remaining constant. 

The nozzles of most commercial types of steam turbines are made 
with straight sides as in Fig. 230, so that only the area at the mouth 
need be determined in addition to that at the throat in order to lay 
out the shape of the tube. 

Equations (157) and (168) are general and are applicable to steam 
of any quality, wet, dry, or superheated. 



STEAM TURBINES 



435 



THEORETICAL DESIGN OF A DIVERGENT NOZZLE 
5000 200 







I'JO 




4500 


180 

iro 




4000 


IGO 




';_) 






^ 


1.50 




O 






3500 !3 


140 


T> 


3 






O" 




Q 


CQ 


IHO 




14 




o 


© 




UJ 


3000°- 


1W 








o 


'C 




p. 


a 


no 








o 


o 




Pm 


2500 a- 


100 


a 


a 






2 


90 


« 


2000 1 


80 


M 


o 




^ 


;i^ 


70 




a 




"^ 


1500 1 


60 


1 


1000 1 


50 
40 




<I 


30 




500 


20 
10 































\ 




























\ 




























\ 
























^y 


^ 


\ 






















^ 


y 




\ 


















y 


/ 










\ 














/ 


X 












\ 










.4^^ 


/ 
















\ 








.;^r 


























/ 




















"-^ 




^. 


/ 


. Q 


Mciii 


ftr 




















/ 


























f 


















"~^ 


■ 








\ 


















y 






/ 




\ 


















/ 




J 


1 






\ 












^/ 


/ 






1 








\ 


s. 






V^^^ 


r^ 


/ 








1 










\ 


^^ 


^ 


U^ 












r- 








_^_ 


-^ 


■— *N 


-^ 


5^^ 


^hr, 










" 
























■*>^ 


r^ 











% 


a 


a 








a 


P^ 


a> 


a, 


!§, 


100^ 


•s 


905 


fcK 




^ 


80 ^ 


o 


>» 


a 

Q 


ro| 



00 



40.000 80,000 120,000 160.000 200,000 240,000 2«0,000 
Kinetic Enei-gy of the Jet, iu Foot Pounds 

Fig. 233. 













■~~ 
















































N, 















__ 


I— 






_4_ 


-->. 




nft 


.06 
















\ 




















^ 


V- 


















K 






















.3 


\ 




^ 
















"^ 


o,*; 


i .05 














\ 


\ 




















\ 




fc 


















V 




















\ 






f^ . 




04 


I-'' 






rvr. 






m 






\ 






p— 

1 


:'M' 


! 1 1 


P. 




\ 




K 






W/y 


















^ |3 










\l 


ort 


,2 -^^ 




f'l- 


VTTT, 


P, 


P- 
1 




p. 














1 










\ 




^ 






8 




















; 1 










fl „^ 






02 


!•" 






1 






2 












p— 

1 


^ 


\ 








^ 










P = 132 Lb. Per Sq. In. Absolute 










I 










.01 


.01 


A 


rea 


ofO 


ritic 


2 0.0 


355 6 


)q.Ii 


'■ 























.1 .2 .3 .4 .5 .0 .7 .8 .'J 1.0 .1 .2 .3 .4 .5 .G .7 .8 .9 1.0 



Ratio— P-^P, 
2 1 



Fig. 234. Flow of Steam through Nozzles. 



436 STEAM POWER PLANT ENGINEERING 

The diameter at the throat may be calculated within an error of 1 to 
2 per cent for the range of pressures usually encountered by means of 
Grashof's formula. 

For dry or wet steam when P„ = < 0.58 Pi 

w' = 60aoPi«-9^ Vi"i. (171) 

For superheated steam when Pn = 0.58 Pi 

w' = 60 ao Pi«-^^ ^ (1 + 0.00065 Q, (172) 

in which 

w' = actual weight of steam discharged, lb. per hr., 

ao = area of the throat, sq. in.. 

Pi = initial absolute pressure, lb. per sq. in., 

Xi = initial quality, 

ts = degree of superheat, deg. fahr. 

For back pressures higher than the critical or Pn> 0.58 Pi the funda- 
mental equation (157) offers the simplest solution. Approximate results 
for this condition may be obtained by multiplying equations (171) and 

(172) by a factor K 

K = 2.182 Vc (1 - 1.19c), (173) 

in which 

C = 1 - (Pn -^ Pi). 

When a divergent nozzle having an actual expansion ratio r ( = mouth 
area -^ throat area) is used for steam pressure having a ratio R ( = 
mouth area -r- throat area for pressure ratio Pn/Pt) a percentage nozzle 
mouth error is introduced of a value Ci = 100 (r — R)^ r, which may 
be positive or negative. Table 87 gives the velocity efficiency or ratio 
of probable actual exit velocity to the theoretical velocity for various 
nozzle mouth errors, assuming the correctly proportioned nozzle to 
have a velocity efficiency of 97 per cent. 

TABLE 87. 

Nozzle mouth error, Ci -40 -30 -20 -10 10 15 20 25 30 

Velocity efficiency, percent 93.5 94.8 95.9 96.7 97 96.7 96.3 95.3 93.6 90.6 

When the actual expansion ratio of the nozzle is greater than re- 
quired, the nozzle is said to be overexpanded; when smaller, under- 
expanded. From Table 87 it appears that it is preferable to have a 
nozzle underexpanded than overexpanded. 

Moyer (''The Steam Turbine," 1st Edition, p. 40) states that the 
ratio of the area of a correctly proportioned nozzle at the throat Aq to 



STEAM TURBINES 



437 



the area at any point a„ is very nearly proportional to the ratio of the 
pressure at point a^ to the initial pressure, or 



an Pn 



(174) 



The entrance to the tube is rounded by any convenient curve. 
The length of the tube may be roughly approximated by the follow- 
ing formula : 

, L = VlK^o, (175) 

in which 

L = length between the throat and mouth, in inches, 
ao = area at the throat, square inches. 

Practice shows that the cross section of a nozzle, whether circular, 
elhptical, square, or rectangular (the latter with rounded corners), has 
very httle influence on the efficiency, provided the inner surfaces are 
smooth and the ratio of the area at the throat to that of the mouth is 
correctly proportioned. The velocity efficiency of a properly propor- 
tioned nozzle with straight sides is about 95 to 97 per cent, correspond- 
ing to an energy efficiency of 92 to 94 per cent, so that it is not considered 
worth while to attempt to follow the more difficult exact curves. 

Example 30. Find the smallest cross section of a frictionless conically 
divergent nozzle for expanding one pound of steam per second from an 
absolute initial pressure of 190 pounds to an absolute back pressure of 
2 pounds and find six intermediate cross sections where the pressures 
will be 70, 30, 14.7, 8, 4, and 2 lb. respectively. Compare the velocity 
and energy of the jet issuing from this nozzle with those of an actual 
nozzle in which 10 per cent of the heat energy is lost in friction. 

From steam and entropy tables we find the values of H, x, u, for 
absolute pressures corresponding to 190, 0.58 X 190 = 110, 70, 30, 
etc., lb. per square inch as follows (theoretical nozzle) : 





H. 


X. 


u. 


S = xu. 


Pi = 190 


1197.3 


1.00 


2.406 


2.406 


P, = 110* 


1152.6 


0.960 


4.047 


3.885 


Pz= 70 


1117.9 


0.932 


6.199 


5.775 


P4 = 30 


1057.2 


0.887 


13.75 


12.27 


P,= 14.7 


1011.3 


0.857 


26.78 


22.95 


P6= 8 


947,8 


0.834 


47.26 


39.29 


P7= 4 


935.6 


0.810 


90.4 


73.2 


Ps= 2 


899.3 


0.788 


173.1 


137.0 



0.58 Pi (= pressure at throat). 



If entropy tables or charts are not available, values Hi to H2, and 
Xi to x% must be calculated. (See Chapter XXII.) 



438 



STEAM POWER PLANT ENGINEERING 



The different quantities for the theoretical nozzle will be calculated 
for the exit pressure Pn = -Ps = 2 lb. per sq. in. absolute. 



d. 



Fs = 223.8 Vh, - H, 



Eg 



As = 



= 223.8 V1197.3 -899.3 
= 3865 feet per second. 

= 778 (^1 - Hs) 

= 778 (1197.3 - 899.3) 

= 232,000 foot-pounds. 

WS 
V 
^ 1 X 137 

3865 
= 0.0353 square foot. 



= ^(lilXi)^. 13.56 VZ 



= 13.56 VO.0353 
= 2.54 inches. 

WVh 



3865 



32.2 
= 120 pounds. 



THEORETICAL NOZZLE 



Quantity | 


V 
Ft. per Sec. 


E 
Fl.-Lb. 


A 

Sq. Ft. 


d 
Inches. 


F 
Pounds. 




(73) 


(72) 


(76c) 




(74) 




Pressures ■ 


110 
70 
30 
14.7 

8 

4 

2 


1,496 
1,995 
2,650 
3,053 
3,339 
3,624 
3,865 


34,767 
61,853 
107,485 
144,742 
173,207 
203,968 
232,000 


.00259 

.00269 

.00461 

.00745 

.0119 

.0202 

.0353 


0.693 

0.702 

0.919 

1.1 

1.46 

1.92 

2.54 


46.4 

61.98 

82.3 

94.8 

103.7 

112.5 

120.0 



In the actual nozzle these values will be modified because of the 
frictional losses. Thus, for Pn = 2 lb., 



78 = 223.8 V {1 - y) {H, - Hs) 

= 223.8 \/(l - 0.1) (1197.3 - 899.3) 
= 3667 ft. per sec. 

Es = 778 (1 - 0.1) (1197.3 - 899.3) = 208,800 ft-lb. 



Xi = 'Xs + h 



STEAM TURBINES 

y (//i - Hs) 



439 



2^8 



^8 



= 0.788 + 



0.1(1197.3-899.3) 



1021 



^8 = 



= 0.788 + 0.029 
= 0.817. 

Wxs'us 



Tom which 



^ 0.817 X 173.1 

3667 
= 0.0386 sq. ft., 

= 2.66 in. 
WVs 3668 



g 



32.2 



= 1141b. 



These various factors for all given pressures have been calculated in 
a similar manner and are as follows: 



ACTUAL NOZZLE. 



. . i 


r 


E 




A 


(1 


P 


Quantities 


Ft. per Sec. 


Ft .-Lb. 


•" • 


Sq. Ft. 


Inches. 


Ft .-Lb. 




110 


1,420 


31,317 


.9658 


.00275 


0.711 


44.1 




70 


1,893 


55,632 


.9414 


.00286 


0.723 


58.8 




30 


2,515 


98,257 


.9026 


.00493 


0.951 


78.12 


Pressures < 


14.7 


2,894 


130,050 


.876 


.0080 


1.2 


98.8 




8 


3,168 


155,858 


.856 


.0127 


1.53 


98.4 




4 


3,438 


183,581 


.836 


.0220 


2.01 


106.8 




2 


3,667 


208,800 


.817 


.0386 


2.66 


114.0 



Many of these values may be determined directly from the Mollier 
or total heat-entropy diagram as described in Chapter XXII; in fact, 
the Mollier diagram has to all intents and purposes supplanted the 
steam tables in this connection. For superheated steam the diagram 
is extremely useful in avoiding laborious calculations. 

Fig. 235 gives a diagrammatic arrangement of the blades in a single- 
stage De Laval turbine. The nozzle directs the steam against the blades 
with absolute velocity Vi and at an angle a with the plane of the wheel 
XX. Since the wheel is moving at a velocity of u feet per second, the 
velocity Vi of the steam relative to the wheel is the resultant of V\ and 
u. The angle ^i between Vi and XX will be the proper blade angle at 
entrance. If the blade curve makes this angle with the direction of 
motion of the wheel no shock will be experienced when the steam enters 
the blades. For convenience in construction the exit angle 02 is made 
the same as the entrance angle fSi. Neglecting frictional losses in the 
blade channels the relative exit velocity will be V2 = V], and the absolute 
velocity V2 is the resultant of V2 and u. The impulse exerted by the 

. . . . . W 

jet in striking the vanes is — V], and its component in the direction of 



440 



STEAM POWER PLANT ENGINEERING 



.WW 
motion is — Vi cos /3i = — (Fi cos a — u). As the jet leaves the vanes 

W W 

the impulse is V2 cos ^2 = {V2 cos y -\- u). 

y y 




Fig. 235. Velocity Diagram. Ideal Single-stage Impulse Turbine. 



The total pressure acting on the vanes, or the actual driving impulse, is 



W 

P = ^ IVi cos a — u — [— (V2 cos y -\- u)]l 

W 

= — (Vi cos a + "^2 COS 7). 



(176) 



Equation (176) may also be expressed 

W 



P = — .2(FiC0sa -u). 



The resultant axial force or end thrust is 

W 



F = — (Fi sin a — V2 sin 7). 



(178) 



(179) 



Evidently if a = 7 and Vi = V2 there will be no end thrust, since 
Vi sin a — V2 sin 7 will be zero. 
The work done is 



W 

Pu = — u (Vi cos a -{- V2 cos 7), 



(180) 



or, using equation (178) in place of (176), 

W 

Pu = — '2u (Fi cos a — u) 

W 

= — '2 {uV\ cos a — u^), 

y 



(181) 



STEAM TURBINES 441 

By making the first derivative equal to zero 

d iW ) 

— j — 2 {uVi cos a — u^)>= Fi cos a — 2 w = 0, 

or u = ^Vi cos a. 

That is, for any nozzle angle a the work done, Pu, has its greatest 
value when w = J Fi cos a or 7 = 90 degrees, whence 

Pu=W-^cos^a. (182) 

The work for any initial velocity Vi becomes a maximum when a = 
and u = i V]. This condition can only occur for a complete reversal of jet 
and zero final velocity. Substitute a = and u = \Vi in equation (181). 

Pu = —^ — , which is necessarily the same as equation (156). 

In the actual turbine the various velocities will be less than those as 
obtained on account of the frictional resistance in the blades, and the 
velocity diagram should be modified accordingly. 

Example 31. Lay out the blades (theoretical and actual) for the 
nozzle in the preceding example, assuming that the jet impinges against 
the wheel at an angle of 20 degrees and that the peripheral velocity is 
1250 feet per second. 

Theoretical Case: 

Lay off Vi = 3865 feet per second in direction and amount as shown 
in Fig. 235 and combine it with u = 1250 feet per second; this gives 
Vi, the relative entrance velocity, as 2725 feet per second, and ^, the' 
entrance angle, as 29 degrees. 

Lay off V2 = Vi at an angle ^2 = (3i and combine with u; this gives 
V2, the absolute exit velocity, as 1740 feet per second. 

The theoretical energy available for doing work is 

W 

= ^ (38652 _ 17402) = 185,000 foot-pounds. 

The difference between 232,000 and 185,000 = 47,000 foot-pounds is 
evidently the kinetic energy lost in the exhaust due to the exit velocity. 
The pressure exerted by the steam on the buckets is 

TF 

P = — ( Fi cos a + F^ cos 7) 

9 
= ^ (3865 X 0.9397 + 1740 X 0.65166) 
= 148 pounds. 
The theoretical impulse efficiency is 

Vi' - yj ^ 38652 _ 17402 ^ 

Y2 38052 "-^y^- 



442 



STEAM POWER PLANT ENGINEERING 



The theoretical horsepower per pound of steam flowing per second is 

_ 185,000 _ 

Theoretical steam consumption per horsepower-hour is 

3600 ,^ ^ 

-^^ = 10.7 pounds. 

Actual Case: 

Proceed as in the theoretical case, using the actual absolute velocity 
Vi = 3865 Vl - y = 3865 Vl - 0.10 = 3667 feet per second in place 
of the theoretical value Fi = 3865. Lay off Vi = 3667 at an angle of 
20 degrees as before and combine with i^ = 1250, Fig. 236. 




— X 



U = 1250 

Fig. 236. Velocity Diagram as Modified by Friction Losses. 

The resultant Vi = 2530 is the velocity of the jet relative to the wheel, 
and the entrance angle /3 is found to be 29.7 degrees. The relative exit 
velocity V2 will be less than Vi because of the blade friction. 

Assume the loss of energy between inlet and exit of the blades to be 
14 per cent; then, since the velocity varies as the square root of the 

energy, 

v^ = Vl Vl - 4> (183) 

= 2530 Vl -0.14 
= 2346 feet per second. 

The resulting absolute velocity V^ is found from the diagram to be 
Vl = 1405 feet per second. 

Since the loss of energy in the nozzle is 



and that in the blade 



Fi^ - (1 - y) 7i^ 

vi' - (1 - 0) Vi^ 

2g 



(184) 
(185) 



STEAM TURBINES 443 

the remaining energy, deducting both losses in the nozzle antl the 
blades, is 

^(y^-yVr-ci>v^-V^) (186) 

= ^ (38652 - 0.1 X 3865^ - 0.14 X 2530^ - UOo^) 
64.4 ^ 

= 164,200. 

The losses due to windage, leakage past the buckets and mechanical 
friction must be deducted from these figures to give the actual energy 
available for doing useful work. Assuming a loss of 15 per cent due to 
this cause, the work delivered is 

0.85 X 164,200 = 139,570 foot-pounds. 

The eflficiency in the ideal case was found to be 0.797 and the avail- 
able energy 185,000 foot-pounds. 

The efficiency, deducting the loss due to friction, etc., is 

139,570 



185,000 
The horsepower delivered is 

139,570 



X 0.797 = 0.60. 



= 254. 



550 

Steam consumption per horsepower-hour is 

3600 ^ , ^ , 

-^^ = 14.2 pounds. 

The heat consumption, B.t.u. per horsepower per minute is 
14.2(1197.3 - 94) 



60 



= 260. 



Assuming the revolutions per minute to be 10,000, the mean diameter 
of the wheel to give a peripheral velocity of 1250 feet per second is 

1250 X 60 o Qo f + oc r • I 

10,000 X 3.14 = 2-^^ ^'''^ ^^ ^^-'' ''''^''' 

The determination of the height and width of vanes, clearance be- 
tween nozzles and blades, etc., are beyond the scope of this work and 
the reader is referred to the accompanying bibliography. 

The ratio of exit to inlet velocity is called the blade or ])ucket velocity 
coefficient. Table 88 gives the values of this coefficient for the usual 
shape of impulse turbine blades. The values include all losses between 
the nozzle mouth and entrance to the exhaust opening. (Marks' Me- 
chanical Engineers' Handbook, p. 984.) 

TABLE 88. 
Velocity relative 

to blades, ft. 

per sec 200 400 600 800 1000 1500 2000 2500 3000 4000 

Blade velocity 

coefficient 0.953 0.918 0.888 0.863 0.811 0.801 0.774 0.754 0.739 0.716 



444 



STEAM POWER PLANT ENGINEERING 



Blade Design for De Laval Turbines: Moyer, "Steam Turbine," Chap, IV; Power, 
Mar. 17, 1908, p. 391. 

Flow of Steam through Nozzles: Jour. A.S.M.E., Mid. Nov., 1909, April, 1910, 
p. 537; Engineering, Feb. 2, 1906; Engr. Lond., Dec. 22, 1905; Eng. Rec, Oct. 26, 
1901; Power, May, 1905; Eng. News, Sept. 19, 1905, p. 204. 

Design of Turbine Disks: Engr. Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481. 

Turbine Losses and their Study: Jour. El. Power and Gas, March 9, 1912. 

Critical Velocity of Shafting: Jour. A.S.M.E., June, 1910, p. 1060; Power, Sept., 
1903, p. 484. 

203. Terry Non-condensing Turbine. — Fig. 237 shows a section 
through a single-stage Terry turbine, illustrating an appUcation of the 
single-stage impulse type with two or more velocity stages. This ''com- 
pounding" of the velocity permits of much lower peripheral velocities 




Fig. 237. Section through Single-stage Terry Steam Turbine. 

than with the single-velocity type. The rotor, a single wheel consisting 
of two steel disks held together by bolts over a steel center, is fitted 
at its periphery with pressed-steel buckets of semi-circular cross section. 
The inner surface of the casing is fitted with a series of gun-metal re- 
versing buckets arranged in groups, each group being supplied with a 
separate nozzle. The steam issuing from nozzle N, at very high velocity, 
Fig. 238, strikes one of the buckets, B, on the wheel, and since the ve- 
locity of the buckets is comparatively low, is reversed in direction 
and directed into the first one of the reversing chambers. The chamber 



STEAM TURBINES 



445 



redirects the jet against the wheel, from which it is again deflected; 
this is repeated four or more times until the available energy has been 
absorbed by the rotor. Terry turbines are made in a number of sizes 
varying from 5 to 800 horsepower, and operate at speeds varying from 




Fig. 238. Arrangement of Buckets and Reversing Chambers in a Terry Steam Tm-bine. 

210 feet per second in the smaller machine to 260 feet per second in the 
larger. These low speed limits compared with the speed of single- 
stage De Laval turbines are made possible by the application of the 
velocity-stage principle in the use of the reversing buckets. The 




Fig. 239. Westinghouse Impulse Turbine Connected to Generator through Re- 
duction Gearing. 

rotor of the smaller machine is 12 inches in diameter and runs at 3800 
r.p.m., and that of the larger, 48 inches, running at 1250 r.p.in. Since 
the flow of steam into and from the buckets is in the plane of the wheel 
there is no end thrust. 

Non-condensing Terry turbines are all of the single-stage type. 




446 STEAM POWER PLANT ENGINEERING 

204. Westinghouse Impulse Turbine. — The Westinghouse impulse tur- 
bine which is constructed in various sizes ranging from 10 to 800 horse- 
power is similar in basic principle to the Terry turbine. The rotor 
consists of a single wheel on the periphery of which are located blades 
of nickel steel. In the non-condensing unit the steam is expanded 
in a single nozzle and is directed upon the rotor where its energy is 
partially absorbed in work. From the rotor it is deflected to the re-, 
versing member and is directed on the wheel a second time when the 

remaining energy is finally extracted. 

For condensing service this reversing 

operation is repeated a second time 

making three passes through the wheel 

before the steam is exhausted to con- 

FiG. 240. Developed Section through denser. By the use of two separate 

Nozzle Blades and Reversing fiozzles, large and small, proportioned 

Chamber, Westinghouse Impulse ^^ g^j^ the load conditions, relatively 

as good efficiency is obtained at half- 
load as at rated load. The economy of a single wheel turbine is vitally 
affected by its operating speed, the higher the speed up to a peripheral 
velocity of less than half that of the jet the better will be the heat 
economy. On the contrary, moderate-speed generators offer a better 
efficiency than high-speed generators. This type of unit has been 
designed so that the steam turbine and its accompanying generator 
may operate at their best speed through the medium of reduction gears. 
In fact, all builders of high-speed turbines are equipped ^to furnish re- 
duction gears with their units, and the general tendency is toward the 
incorporation of gearing in all types under 1000 kw. rated capacity. 

205. Elementary Theory. — Single-wheel Multi-velocity-stage Turbine. — 
Fig. 241 gives the theoretical velocity diagram for a single-pressure stage 
Terry Turbine. Since the entire heat drop takes place in the nozzle 
the initial velocity of the jet OA is the same as with the single-stage 
De Laval turbine and may be calculated by means of equation (157). 
OA represents the absolute velocity of the jet, OC the peripheral velocity 
and AOC the angle of the nozzle. CB is the component, parallel to 
the line of the jet, of the resultant of AO and OC. DC, in line with 
and equal in length to CB, combined with the peripheral velocity DE 
gives EC, the absolute velocity of the steam as it leaves the first set of 
rotating buckets. OiF, parallel to OA and equal in length to EC, 
represents the velocity of the steam as it enters the first stationary or 
reversing bucket. JG is the component of the resultant of OiF and 
OiJ in line with the jet. The resultant // of HJ (= JG) and HI 
represents the velocity of the steam as it leaves the rotary buckets the 



STEAM TURBINES 



447 



(oa' - wy'). 



second time. This construction is repeated through all velocity stages. 
The final exit velocity of the steam as it issues from the moving buckets 
is WY. The energy converted into useful work is 

In the actual turbine friction losses would reduce the length of the 
velocity lines and increase the amount of energy rejected in the exhaust. 
The construction of the velocity diagram as modified by friction is 
similar to that described in paragraph 202. Fig. 235. 




Fig. 241. Theoretical Velocity Diagram, Terry Turbine. 

206. De Laval Velocity-stage Turbine. — Fig. 242 shows a section through 
a ''Class C" De Laval steam turbine illustrating an application of the 
single-pressure, multi-velocity stage type in which the velocity is com- 
pounded by a series of wheels and reversing intermediates instead of 
having the jet redirected upon a single rotor. In this type of turbine 
the steam is completely expanded in a single set of nozzles from initial 
to terminal pressure just as in the single- wheel geared type. The jet 
from the nozzles impinges against the first row of moving blades or 
vanes and gives up part of its energy. It leaves the moving blades at a 
reduced velocity and is reversed in direction by the first set of station- 
ary vanes. The latter redirect the jet against the second set of moving 
vanes where a further absorption of energy takes place and the velocity 
is again lowered. This process is repeated until the steam leaves the 
last row of moving vanes at practically zero velocity. The wheels are 
of forged steel and are fitted at the periphery with nickel l^ronze blades 
similar in design to those of the single-stage geared type. The guide 
vanes are similar in form to the moving vanes and are attached in a 
Hke manner to a steel retaining ring. The governor is of the throttling 
type. In the smallest machine the governor weights are attached di- 



448 



STEAM POWER PLANT ENGINEERING 



rectly to the main shaft and in the larger machines it is actuated through 
speed reduction gearing. The emergency governor is independent of 
the main governor and closes a butterfly valve in the steam inlet open- 
ing when a predetermined speed is exceeded. This type of turbine is 




Fig. 242. De Laval Velocity-stage Turbine. 

constructed in sizes ranging from 1 to 1500 horsepower and at speeds 
ranging from 3600 to 6000 r.p.m. By means of reduction gearing any 
desired lower speed may be obtained. 

The velocity diagram may be constructed in a manner similar to 
that of any single-pressure stage of the Curtis turbine as described in 
paragraph 212. 

307. Kerr Turbine. — Fig. 243 shows a longitudinal section through 
an eight-stage Kerr steam turbine illustrating the compound-pressure 
or multi-cellular group of the impulse type. The rotor consists of a 
series of steel disks, mounted on a rigid steel shaft. A series of drop- 
forged steel buckets is secured to the periphery and riveted in dove- 
tailed slots as shown in Fig. 244. The tips of the buckets are riveted 
to a shroud ring, thereby insuring a rigid and positive spaced construc- 
tion. The stator is made up of a number of arched cast-iron diaphragms 
with circular rims tongued and grooved, and bolted to steam-end and 
exhaust-end castings. The nozzles are formed by walls within the 
diaphragm and thin Monel metal vanes die-pressed into shape and cast 



STEAM TURBINES 



449 







bC 



bO 

o 



450 



STEAM POWER PLANT ENGINEERING 



into the diaphragm. One set of nozzles and one wheel constitute a 

stage and the expansion is usually carried out in from six to ten stages, 

. -,„ -^ _; ^ ^ depending upon the condition 




Fig. 244. Bucket Fastening, Kerr Turbine. 



of operation. 

The operation is as fol- 
lows: Steam enters the tur- 
bine through a double-beat 
balanced poppet valve, the 
stem of which is connected 
through levers to the gover- 
nor, to the circular cored 
space H, H extending around 
the steam ''end casting." 
This space acts as an equal- 
izer and insures uniform ad- 
mission to the first set of 
nozzles. Partial expansion 
takes place through the first 



set of nozzles and the kinetic energy is imparted to the rotor through 

the medium of the vanes. 

Steam leaves the buckets at a 

very low velocity and is again 

expanded through the second 

set of nozzles in the diaphragm. 

This process is repeated in each 

stage and exhaust steam leaves 

the turbine at 0. 

Fig. 246 illustrates the prin- 
ciples of the oil relay governor 
as applied to the larger sizes of 
turbines driving alternators. 
Referring to Fig. 246: rotation 
of the turbine shaft is trans- 
mitted through worm gear and 
governor spindle to weights, W, 
W. Centrifugal force throws 
these weights outward about 
suspension points A and A\ 
overcoming the resistance of 
the spring. The movement of 
the spring is transmitted through lever L to relay plunger P and ad- 
mits oil pressure (about 30 pounds per square inch) to piston *S and 




Fig. 



245. Arrangement of Vanes and 
Nozzles, Kerr Turbine. 



STEAM TURBINES 



451 



in this manner throttles admission valve V. Similarly, a downward 
movement of the relay plunger stem releases oil pressure and opens 
the admission valve. 

Floating lever L is connected to the admission valve stem through 
secondary lever M so that the movement of the steam valve returns 




Fig. 246. Oil Relay Governor, Kerr Turbine. 

the relay plunger to its central position. This equalizes the pressure on 
top and bottom of the main piston S and arrests its movement, thereby 
maintaining a fixed opening for a given speed. A suitable emergency 
valve automatically cuts off the steam supply in case the speed exceeds 
a predetermined amount. 

A spring-loaded governor of the centrifugal type mounted directly 
on the turbine shaft is used to control the smaller sizes of turbines. 

Kerr turbines are constructed horizontally and vertically and in 
various sizes ranging from 5 to 2500 horsepower, and are designed to 
operate all classes of pumps, blowers and generators. The rotative speed 
varies from 2000 to 4000 r.p.m., depending upon the service for which 
the turbines are intended. By means of gearing any lower speed may 
be obtained. 

208. De Laval Multi-stage Turbine. — This is of the multi-cellular type 
and is constructed with single velocity stages or with two velocity 
stages for each pressure stage. The increase in the cross-sectional area 
of the passages required by the expansion of the steam as it proceeds 



452 STEAM POWER PLANT ENGINEERING 

through the turbine is effected by lengthening the blades, reducing the 
diameters of the wheels correspondingly and increasing the bore of the 
casing. (In the Kerr turbine the blades are lengthened and increased 
in width from the high-pressure to the low-pressure stages and the steam 
passages are increased in size but the outside diameter of the rotor 
remains the same.) The bearings are of rigid construction arranged 
for water cooling. Labyrinth packing is used between stages and com- 
bined labyrinth and carbon-ring packing at the steam and exhaust 
ends of the casing. Air leakage into the turbine is prevented by in- 
troducing live steam between the two outer carbon rings. The governor 
is of the throttling type and is mounted upon a vertical shaft driven 
through worm gearing by the main turbine shaft. These machines 
are constructed in various sizes ranging from 50 to 15,000 horsepower. 
The maximum speed of the smaller machines is about 7500 r.p.m. 

209. Elementary Theory. — Multi-pressure Single-velocity-stage Turbine. 
In the frictionless or ideal turbine the velocity issuing from each nozzle 
or pressure stage is dependent upon the heat drop in the nozzle. If 

there are n stages the heat drop per stage will be - of the total heat 

drop. Since there are no friction losses in the ideal turbine the total 
heat drop is H — H 

and the heat drop per stage „ ^j 



The stage velocity or initial velocity of jet from each nozzle is 
V = 224 



V n 



The pressure, specific volume and quality of the steam in each stage 

IT JJ 

may be determined by subtracting — from the heat content of 

the preceding stg^ge and finding the corresponding quantities from tem- 
perature-entropy tables or diagrams. 

Thus, an eight-stage turbine operating non-condensing at 190 pounds 
initial absolute pressure would show about the following conditions. 
(All friction and leakage losses neglected and final velocity in each 
stage assumed to be zero.) 

i7i = 1197.3 B.t.u. per pound. 

Hn = 1012.5 B.t.u. per pound. 

Total heat drop = Hi-Hn = 1197.3 - 1012.5 = 184.8. 

Heat drop per stage = ' = 23.1. 

o 

stage velocity = 224 \/23JL = 1080 feet per second. 



STEAM TURBINES 



453 



Stage. 


Heat Content. 


Pressure, Lb. Abs. 


Quality, Per Cent. 


Specific Volume 
Cu. Ft. per Lb. 


Admission. 


1197.3 


190 


100 


2.41 


1 


1174.2 


145 


97.9 


3.04 


2 


1151.1 


109 


95.9 


3.93 


3 


1128,0 


80 


94.0 


5.14 


4 


1104.9 


58 


92.2 


6.77 


5 


1081.8 


42 


89.6 


8.96 


6 


1058.7 


30 


88.8 


12.07 


7 


1035.6 


21 


87.3 


16.33 


8 


1012.5 


Atmospheric 


85.8 


22.55 



In the actual turbine only 50 to 75 per cent of the heat theoretically 
available is transformed into useful work. A small portion is lost by 
gland leakage, radiation and bearing friction and the balance has been 
retransformed from kinetic energy into potential energy by eddying, 
fluid friction and blade leakage. The efficiency of each stage is less 
than that of the turbine as a whole since the increase in heat content 
due to friction, etc., is available for transformation into useful work in 
the succeeding stages. To find the actual pressure condition in each 
stage allowing for the various losses, it is necessary to correct the theo- 
retical quantities for these losses. See '' Energy and Pressure Drop in 
Compound Steam Turbines," by F. E. Cardullo, Proc. A.S.M.E., Feb., 
1911, and paper read by Prof. C. H. Peabody, Proc. Society of Naval 
Architects and Marine Engineers, June, 1909. Consult also, ''The 
Steam Turbine Expansion Line on the MoUier Diagram and a Short 
Method of Finding the Reheat Factor," by Edgar Buckingham, Bui. 
No. 167, 1911, U. S. Bureau of Standards. 

210. Terry Condensing Turbine. — The condensing units are of a 
composite design, namely, a high-pressure compound-velocity element 
similar to the non-condensing device and a series of single-velocity multi- 
pressure elements for the low-pressure end. A section through such a 
unit is shown in Fig. 247. It will be noted that the steam (slightly 
above atmospheric pressure) leaving the high-pressure element instead 
of passing directly into the low-pressure stages passes to the other end 
of the casing and returns through the low-pressure stages to the center. 
This arrangement maintains a pressure somewhat above atmospheric 
on the inside of both glands and prevents inward leakage of air. It 
also insures uniform temperature at both ends of the turbine casing. 
These units are built in sizes up to 750 kw. 

211. Curtis Turbine. — A textbook description of the Curtis line of 
turbines must necessarily be of a very general nature because of the 
many changes effected from year to year. A detailed description of a 



454 



STEAM POWER PLANT ENGINEERING 




Fig. 247. Section through Terry Condensing Steam Turbine. 




IJozzle Ports 



NExLaust Connection 

Fig. 248. Curtis Single-stage Turbine. 



STEAM TURBINES 



455 



machine representing current practice may be obsolete in many respects 
when compared with that of a similar unit constructed a year later. In 
a general sense the basic principle is the same for all types and sizes, but 
the structural details, methods of governing, number of stages, and the 
like vary with the size and the service for which the turbine is intended. 
All Curtis turbines ranging from the very small direct-current machine to 
the huge turbo-alternator of 45,000 kilowatts rated capacity are of the 
impulse type. The small direct-current machines, Fig. 248, have a 
single-pressure stage, and two or three velocity stages and operate at 
approximately 5000 r.p.m. The large turbo-alternators have nine or 



Steam Admission 




Fig. 249. Arrangement of Nozzle and Blades, Curtis Turbine. 



more pressure stages contained in cither a single cylinder casing, Fig. 
251 or double cyUnder casing, Fig. 256 and operate at 1800 r.p.m. 
The high-pressure stage in practically all sizes comprises a set of nozzles 
and a single wheel carrying two rows of buckets. The succeeding 
stages have but one row of buckets on a single wheel, except in the low- 
pressure element of the compound cylinder units where there are two 
wheels per stage, each with a single row of buckets. In the mixed 
pressure type there are two rows of buckets on each wheel and in the 
small, two-stage turbo-oil pump for circulating the main bearing oil 
there is but one velocity for each pressure stage. In all machines the 
steam flow is axial. In the single cylinder units the flow is unidirectional 
but the low-pressure member of the compound cylinder machine is 



456 



STEAM POWER PLANT ENGINEERING 



arranged for double flow, that is, the steam enters at the center of the 

low-pressure turbine and flows through double stages on either side of 

the condenser as shown in Fig. 251. 

A comparatively high initial velocity is given to the jet in each 

pressure stage by expansion in the nozzle and the energy is absorbed 

by successive action upon a 
series of moving and station- 
ary vanes. Since expansion 
takes place only in the nozzles 
the pressure in any stage is 
the same on both sides of the 
wheel. The action of the 
steam is as follows (Fig. 249): 
Entering at A from the steam 
pipe it passes through one or 
more admission valves B into 
the bowls C. The number of 
leei admission valves depends on 
,fi the load and their action is 
controlled by the governor. 
From bowls C the steam ex- 
pands through nozzles D and 
impinges against the first row 
of moving blades and gives up 

Fig. 250. Throttling Governor Mechanism for part of its energy. The steam 
25-kw., 3600-r.p.m., Direct-current Curtis flowing from the first row of 
Turbo-Generator. ^^^.^^ ^j^^^^ .^ reversed in 

direction by the adjacent stationary vanes and is redirected against the 
second set of moving blades where it gives up its remaining kinetic 
energy. From this stage the steam flows at reduced pressure through 
the nozzles of the second stage which are sufficient in number and 
size to afford the greater area required by increased volume. In ex- 
panding in these nozzles it acquires new velocity and gives up energy 
to the moving blades as before. This process is repeated through 
several additional stages. 

The rotor consists of 1 to 13 or more steel disks mounted side by side 
on a horizontal shaft. In some of the earher designs the shaft was 
mounted vertically but this construction has been discontinued. 
Buckets or vanes of nickel steel, monel metal or nickel bronze, ac- 
cording to the condition of the steam, are secured to the periphery by a 
dovetail-shaped root which fits snugly in a channel of the same section 
machined in the rim. The types of the vanes are tenoned and riveted 




STEM! TUP JUNES 



457 




458 



STEAM POWER PLANT ENGINEERING 



into a shroud ring. The stationary vanes are secured to the casing as 
illustrated in Fig. 249. Between the revolving wheels is a stationary 
steam-tight diaphragm which contains the nozzles through which the 
steam is expanded from the preceding stage. It will be seen from Fig. 
249 that vanes and nozzles increase in size in succeeding stages as the 



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012345G78 
Steam Belt Area 

Fig. 252. Steam-belt Area in Five-stage Curtis Turbine. 

pressure falls and the volume increases. The parts are so proportioned 
that the steam gives up approximately \/n of its energy in each pressure 
stage, n representing the number of stages. The number of stages 
and the number of vanes in each stage are governed by the degree of 
expansion, the peripheral velocity which is practical or desirable, and 
by various conditions of mechanical expediency. The nozzles extend 




Fig. 253. Main Operating Governor. 

around a relatively short arc in the periphery of the first stage and in- 
crease progressively in number until they extend around the entire 
wheel in the last stage. See Fig. 252. 

In the smaller machines the speed is controlled by a centrifugal gov- 
ernor mounted on the end of the main shaft. The governor actuates 
a throttling valve of the balanced poppet-valve type. The larger 
sizes are controlled by an indirect or relay system. Fig. 250 shows an 



STEAM TURBINES 



459 



assembly of the simple throttling governor. The movement of gov- 
ernor weights B is transmitted through spindle C and thrust ball D 
to bell crank E, which in turn operates the valve stem F and double 
balanced poppet valve G. The valve is shown in wide open position. 
Two emergency governors of the clock-spring type are mounted on the 
outer ring of the main governor. These springs rest under tension 
against a stop and when the turbine exceeds 10 per cent of the normal 





Fig. 254. Assembly of Hydraulic 
Cylinder. 



Fig. 255. Main Controlling 
Governor. 



speed will strike a trigger and release flap valve W, thus shutting off 
the supply of steam. 

The large turbo-alternators are controlled by a relay governor of the 
hydraulic type. This mechanism consists of a cylinder to which oil 
under pressure is fed through a pilot valve under the control of the main 
governor. The piston rod of the cylinder contains a rack which meshes 
with a pinion on a cam shaft and so rotates the shaft. The cams on 
the shaft lift the individual controlling valves as determined by the 
angular spacing of the cams. The general assembly is shown in Fig. 
251 and the general details of the main governor, cylinder and one of 
the controlling valves in Figs. 253 to 255 respectively. 

Referring to Fig. 253 speed regulation is accomplished by the balance 
maintained between the centrifugal force of moving weights A A and 
the static force exerted by spring D. The governor is provided with 
an auxiUary spring F for varying its speed when synchronizing, the 



460 



STEAM POWER PLANT ENGINEERING 




tension of which is varied by 
a small pilot motor controlled 
from the switchboard. The 
movement of the governor 
weight is transmitted through 
rod C to arm H and by means 
of the latter to floating lever L, 
Fig. 254. This floating lever is 
pivoted on a clamp attached to 
the pilot valve stem S. The 
other end of the lever is con- 
nected by links to the piston rod 
R of the operating cyhnder. A 
movement of governor arm dis- 
places the small pistons of the 
pilot valve from their normal 
location in which they close the 
ports of the cyhnder. This dis- 
placement causes oil to be ad- 
mitted to the cylinder and the 
pressure of the oil operates the 
main piston. The piston rod 
opens and closes the controlhng 
valves through the agency of 
the cam shaft and at the same 
time transmits its motion 
through the link system to the 
end of the floating lever and 
thus brings the pilot valve back 
to its normal position. Each 
position of the governor deter- 
mines a definite position of the 
piston in the operating cylinder 
and consequently the opening of 
a definite number of controlling 
valves. The general details of 
one of the controlling valves is 
shown in Fig. 255. 

The emergency governor or 
stop consists of a ring R (Fig. 
257), unevenly weighted, at- 
tached to and revolving with 



STEAM TURBINES 



461 



the shaft. At normal speeds and less, the unbalanced ring is held con- 
centric with the shaft by helical springs S. When the speed increases 
to 10 per cent above normal the centrifugal force of the unbalanced 
portion of the ring overcomes the spring tension, and the ring revolves 
eccentrically. In this position the ring 
strikes a trip finger and closes the main 
throttle valve which is of the balanced 
type. 

Another type of governor used on 
the smaller machines is shown assem- 
bled in Fig. 258. In this arrangement 
the main governor arm A, Fig. 259, 
actuates a small steam pilot valve. 
The latter in turn moves a piston on 
the stem of which are mounted several valve hangers designed to raise 
the various controlling valves successively with the upward travel of 
the piston. The valves are of the double-seated poppet type, an- 
nular in shape, free to move, but guided and controlled by the valve 
hangers. Referring to Fig. 259 with turbine and governor at rest 
and no steam bled to the pilot valve, the position of the various 




Fig. 257. Emergency Governor. 




Fig. 2.^8. Assembly of Governor and Operating Cylinders for Steam Relay Control. 

levers is such that the pilot valve is in a position to admit steam to 
the under side of the piston for operating the main valves. With the 
opening of the throttle the tur})ine speeds up and the governor mecha- 
nism moves upward, the connection to floating lever L moves down- 



i62 



STEAM POWER PLANT ENGINEERING 



ward and the latter fulcruming on pin B moves the pilot valve V 
upward to a position which shifts the admission of steam from the 
under to the upper side of piston P, closing the main valves successively 
until the governor assumes a position of equiUbrium. Movement of 



Sleam Inlet 
Steam Exhaust 




Fig. 259. Valve Gear Assembly. 

piston P also moves floating lever L and brings the pilot valve in a 
neutral position independent of the governor. Any change in speed of 
the governor causes the pilot valve to admit steam to the under side 
of the piston with a drop in speed and the upper side with an increase 
in speed. It will be seen from the above description that throttling 



STEAM TURBINES 



463 



is practically eliminated since all but one of the valves is either in the 
wide open or shut position, and the over travel of the piston during a 





Cooling Coil 

/Ul LJ 



m 



(q° b"cj izj CD a a 





Oil Feed Inlefr 



Fig. 260. Water-cooled Bearing, Horizontal Curtis Turbine. 

small variation in load causes periodic opening and closing of the in- 
dividual valve. 

The main bearings as well as the governor are supplied with forced 
lubrication from an oil pump 



Thrust shell 




Outer 

thrust shell 



bolted to the main pillow block. 
The oil is cooled by a current of 
water flowing through the main 
bearing linings, the bottom 
halves of which are equipped 
with a number of copper coils 
imbedded in the babbitt as 
shown in Fig. 260. In some 
designs the oil pump is of the 
geared type while in others it 
is driven by a small independ- 
ently operated steam turbine. 

The general details of the 
main thrust bearing are shown 
in Fig. 259. The drawing is 
self-explanatory. 

Fig. 262 shows a general assembly of the main shaft packing for the 
ends of the casing and for one intermediate. The packing rings are 
of carbon and are self-lubricating. Each ring is composed of three 



Fig. 261. Details of Thrust Bearing. 



464 



STEAM POWER PLANT ENGINEERING 



segments held against the shaft by radial springs. There are usually 
four rings in the high-pressure head, three in the low-pressure head 
and a single ring in the intermediates. For condensing service the 
heads and packing chambers are sealed with live steam to prevent 



Live Steam Admission 



Diaphragm 
Steam Seal Inlets 




Packing Box 
Drains 



Fig. 262. Assembly of Packing. 



air leakage into the casing when the pressure inside is less than atmos- 
pheric and to prevent the escape of steam when the pressure in the 
casing is greater than atmospheric. 

Although there are numerous Curtis turbines of the Curtis vertical 
.shaft type in successful operation this design has been discontinued 
and no attempt will be made to describe them. The mechanical valve 
employed in some of the earlier designs has also been abandoned. 

312. Elementary Theory, Curtis Turbine. — Fig. 263 gives a dia- 
grammatic arrangement of the blades and nozzles in the first stage 
of a two-stage Curtis turbine, each stage consisting of one set of nozzles 
and two moving and one stationary sets of blades. 

Referring to the diagram: the steam is expanded in the first stage 
from pressure Pi to P2 and issues from the first set of nozzles with 
absolute velocity Fi, striking the first set of moving blades at an angle 
a with the line of motion of the wheel. The resultant Vi of Vi and the 
peripheral velocity u is the velocity of the steam relative to the vanes; 
and the angle jS which the Une Vi makes with the hne of motion of the 
wheel is the proper entrance angle of the blades for the first set. Neg- 
lecting friction the exit angle 7 will be the same as the entrance angle 
/3. The resultant of Vi, the exit velocity relative to the blade, and w, 
the peripheral velocity, is F2, the absolute exit velocity. 



STEAM TURBINES 



465 




Fig. 263. 



Since the second set of blades is fixed and serves as a means of 
changing the direction of flow, the absolute velocity entering them is 
V2. The angle 8 formed })y V2 and the center line of the stationary 
blades is the proper entrance angle. Neglecting friction the absolute 
exit velocity will be V-^ = V2, and the 
exit angle will be e = 5. The steam 
flowing from the stationary blades 
strikes the second set of moving 
blades at an angle e = 8 with absolute 
velocity V3. Combining V^ with the 
peripheral velocity u we get v^, the 
velocity of the steam relative to 
the second set of moving blades. 
The angle 6, formed by v^, and the 
hne of motion of the wheel, is the 
proper entrance angle for the second 
set of moving blades. The resultant 
of z;4 ( = Vz) and u is F4, the absolute 
exit velocity for the first stage. 

In the second stage the steam is 
expanded from pressure P2 to that 

in the condenser and acquires initial velocity Va, leaving the last bucket 
with residual velocity F„. The theoretical velocities and blade angles 
for this stage may be found as above. 

Example 32. A four-stage Curtis turbine develops 800 horsepower 
on a steam consumption of 12 pounds per horsepower-hour; initial 
pressure 150 pounds absolute, superheat 100 deg. fahr., back pressure 
1.5 pounds absolute, peripheral velocity 450 feet per second, angle of 
the nozzle with the plane of rotation, 20 degrees. Each stage consists 
of two rotating elements and one stationary element. Compare the 
performance of the actual turbine with its theoretical possibilities. 

Ideal Turbine: 

For the sake of simplicity it will be assumed that the final velocity 
of each stage is zero and that the heat drop in the first set of nozzles is 
one-fourth of the total theoretical drop assuming adiabatic expansion. 
From steam tables Hi = 1249.6 B.t.u. 
From entropy tables or Mollier diagram Hn = 934.6. 
Total heat drop = 1249.6 - 934.6 = 315. 
Heat drop in first stage ^l^ = 78.75. 
The velocity of the jet in the first stage is 



Velocity Diagram, Curtis 
Turbine. 



Vi = 224 V78.75 = 1985 feet per second. 
By laying off this initial velocity in direction and amount and com- 



466 STEAM POWER PLANT ENGINEERING 

bining it with the peripheral velocity as in Fig 263, the absolute veloci- 
ties V2 and V's may be readily obtained. 

The kinetic energy absorbed in the first set of moving blades, per 
pound of steam, is 

= ^ (19852 - 11702) = 39,930 foot-pounds per second, 
and in the second set of moving blades 

= 14,280 foot-pounds per second. 
The total energy converted into useful work is 

39,930 + 14,280 = 54,210 foot-pounds per second. 
Had the entire heat drop been utilized in doing work the total energy 
would be 

wT-j X 19852 = 61,180 foot-pounds per second. 

The difference 61,180 - 54,210 = 6970 represents the loss due to the 
residual velocity of the steam leaving the last bucket. 

Since the steam is brought to rest before entering the second set of 

nozzles, the heat equivalent of this energy or = 8.96 B.t.u. in- 

77o 

creases the final heat content; thus 

H.2 = 1249.6 - 78.75 + 8.96 = 1179.8 B.t.u. 

But a total heat drop per stage of 78.75 B.t.u. was assumed as a 
requirement of the problem and the final result obtained above shows 
it to be 78.5 — 8.96 = 69.54. By trial and adjustment or by means of 
empirical formulas a value of Hy may be obtained which will fulfill the 
given conditions. Such an analysis is beyond the scope of this book, 
and the reader is referred to Forrest E. Cardullo's article ''Energy and 
Pressure Drops in Compound Steam Turbines," Trans. A.S.M.E., 
vol. 33, p. 325, 1911. 

The remaining stages may be analyzed in a similar manner. 

It should be borne in mind that in the actual turbine the velocity 
will be less than the theoretical on account of frictional resistances in 
the nozzles and blades and the heat content Hi, H2 . . . Hn will be 
greater than that of the ideal mechanism. Radiation, leakage, windage 
and other losses must also be considered in determining actual conditions. 



STEAM TURBINES 407 

Neglecting the residual energy in the exhaust, the total heat drop 
Hi — Hn is available for doing useful work and the water rate of the 
ideal turbine is 

2546 2546 

W = r- — = -—— = 8.1 pounds per horsepower-hour. 

Heat consumption per horsepower per minute 

8.1 (1249.6 - 83.9) ^ „ _. ^ 

= ^7, = lo7 B.t.u. 

60 

Thermal efficiency 

_ 1249.6 - 934.6 

^' ~ 1249.6 - 83.9 ~ ^''^^' 

Actual Turbine: 

Steam used per hour = 800 X 12 = 9600 pounds. 

Steam used per second = 9600 -^ 3600 = 2.66 pounds. 

Horsepower developed per pound of steam flowing per second = 

800 -- 2.66 = 300. 

Kinetic energy converted into useful work: 

300 X 550 = 165,000 foot-pounds per second. 

Thermal efficiencv 

p _ 2546 ^ 

' 12(1249.6-83.9) 

Heat consumption, B.t.u. per horsepower per minute, 

12 (1249.6 - 83.9) 



60 



= 233. 



Rankine cycle ratio = -~ = ^' _„ = 0.675. 
rLr U.Zn) 

213. Westinghouse Single-flow Reaction Turbine. — The Westinghouse- 
Parsons single-flow reaction turbine was one of the first reaction tur- 
bines in successful use in this country. This particular type of 
machine is no longer constructed, though the modern Westinghouse 
non-condensing reaction turbine and the high-pressure element of the 
large compound units are modifications of the conventional Parson 
design. The reaction turbine is always a multi-pressure-stage machine 
with small pressure drop per stage. Each stage consists of a stationary 
set of blades or nozzles and a row of rotating vanes or buckets. The 
stationary blades are inserted radially into circumferential grooves in 
the main cylinder or are carried in separate blade rings which are 
centered within the cylinder. The rotating vanes are mounted in 
rows on a steel barrel or drum and when in operating position revolve 
between the rows of fixed blades on the stator. Theoretically each 



!</' 



468 STEAM POWER PLANT ENGINEERING 

set of stationary and moving buckets should either continually in- 
crease in height from one end to the other to accommodate the increas- 
ing volume of steam, or, with equal heights of blades the equivalent 
nozzle area should gradually increase. Practically, for constructive 
reasons, it is preferable to subdivide the stages into three divisions, 
each with a different pitch diameter, and to arrange each division into 
a number of groups. The blades in each group are of the same height 
and shape but are so assembled or ''gagged" that the equivalent nozzle 
area gradually increases as the steam volume increases. The entire 
expansion is effected in the annular compartment between rotor and 
stator and resembles in effect a single divergent nozzle with the ex- 
ception that the dynamic relationship of jet and vane is such as to 
secure a comparatively low velocity from inlet to outlet. The action 
of the steam on the blades is illustrated in Fig. 264. Steam is expanded 
p/ , in the first row of stationary blades 

C ( C ( C\( C C C^tationiry Blades from prcssure P to Pi and accel- 
)) i) ))J% )) )) ^ )) 4 ^^l^ erates the jet. The velocity of the 

CC C( C( C CCst'^tioniry Blades jct issulug from thcse stationary 
J) J) J) O ]) J) D F( ^'gfca nozzles is such that steam enters 
^ ^^ the adjacent set of moving blades 

Fig. 264. Blade Arrangement, Reaction practically without impulse. The 
lurbme. steam expands from pressure Pi 

to P2 in passing through the first set of moving blades and exerts a 
reactive force on the blades. The jet, with low residual velocity, is 
deflected from the moving blades to the entrance of the second set of 
stationary nozzles. In this second set of stationary nozzles the steam 
is expanded from pressure Po to P3 and the jet strikes the second set 
of moving blades. This process is repeated in each element of the 
turbine, the steam expanding as it flows from element to element in 
its passage to the condenser. It will be seen that the rotating force 
is primarily due to reaction though there may be some impulse when 
the jet strikes the moving members. Since all reaction turbines are 
subject to an axial end thrust of the rotating parts due to the differ- 
ence of steam pressures at each end of the drum, provision must be 
made for resisting this thrust. In the original Westinghouse-Parsons 
turbine this was effected by balancing pistons or ''dummies" mounted 
on the rotor, running with close clearance to the casing, and of the 
same diameter as the overall diameter of each drum. Each dummy is 
then subjected to the same difference of pressure as the rotating drums 
by means of equalizing pipes. In the modern single-flow design there 
is but one balancing piston and the thrust is taken up by a suitably 
designed thrust bearing. A section through a modern Westinghouse 



STEAM TURBINES 



469 




:3 
o 



OS 

I 



470 



STEAM POWER PLANT ENGINEERING 



non-condensing reaction unit is illustrated in Fig. 265. It will be 
seen that grooves are cut in the surface of the dummy in which cor- 
firHrhr^sT==i responding collars on the turbine 

^' 111'^ ™i ' " ''''' ' mesh, although they run without 

actual metalhc contact. The dummy 

is split and the openings lead to the 

interior of the spindle. Thus the 

steam leaking past the inner half of 

the balancing piston is conducted 

through the spindle to the low- 

FiG. 266. Method of Fastening pressure stages and does work in 

Reaction Blading. the low-pressure cylinder. An equi- 

hbrium pipe connects the chamber housing the balance piston with 

the exhaust chamber and serves to equalize the pressure at both ends 





Fig. 267. Assembly of Governor Mechanism, Oil Relay Control. 

of the spindle and at the same time permits steam leakage past the 
outer half of the piston to escape into the exhaust. 

Fig. 266 shows the method of fastening the roots of the blades and 



STEAM TURBINES 



471 



of bracing the upper ends. It will be seen that dovetailed packing 
pieces placed between the blades and on top of the upset root provide 
an interlocking system with which no caulking is necessary. This 
arrangement makes it possible to replace the blade without mutilating 
the blade-carrying member. 

Fig. 267 shows an assembly of the main governor mechanism. The 
movement of the governor weights is transmitted through suitable 
linkage to lever L which in turn actuates rocker R. Flat-faced cam C 
and vibrator rod B impart a slight but continuous reciprocating motion 
to lever L and thus overcome the friction of rest. Rocker R controls 




Fig. 268. Oil Relay Valvo Gear. 



a small pilot valve T" which admits oil under pressure to or exhausts 
it from the admission valve operating cylinder. Fig. 268 shows the 
general details of the oil relay gear. Relay valve A is controlled by 
the governor and admits or exhausts oil from the operating cylinders. 
When oil is admitted to the operating cyhnder raising the piston, the 
lever C lifts the primary valve E. The lever D moves simultaneously 
with C but on account of the slotted connection with the steam of the 
secondary valve F the latter does not begin to lift until the primary 
valve is raised to the point at which its effective opening ceases to be 
increased by further upward travel. The secondary valve admits 
steam to the intermediate section and enables the turbine to carry 
about 50 per cent more load than on the primary valve alone. A 
steam relay gear is also used with this type of turbine. The steam 



472 



STEAM POWER PLANT ENGINEERING 



chest in this case contains only one valve — the primary valve — 
which is operated by the steam relay gear. All capacities in excess 
of that of the primary valve are carried by means of a hand-operated 
secondary valve. All turbines are equipped with an automatic gov- 
ernor stop which shuts off the main steam supply when the turbine 
speed exceeds a predetermined limit. 

In the smaller sized machine running at 3000 r.p.m. or more, flexible 
bearings are employed to absorb the vibration incident to the critical 




Fig. 269. Direct Control Mechanism for High-pressure Turbine Steam- 
admission Valve. 

velocity. They consist of a nest of loosely fitting concentric bronze 
sleeves with sufficient clearance between them to insure the formation 
of a film of oil. In the larger and lower speed machines a split self- 
aligning babbitted bearing is used instead of the flexible bearing. 
Before the babbitt is run in, a large copper tube is placed in a groove 
cast in the shell. This tube receives the oil and deh vers it to the top 
of the bearing. 

A closed oiUng system is maintained by means of a pump geared 
to the main shaft of the turbine. The oil, after it drains from the 



STEAM TURBINES 



473 



bearing, passes through a strainer into a collecting reservoir whence 
it is pumped through a cooler and back to the bearings. In turbines 
in which the oil-relay 
governing system is em- 
ployed, and a higher 
pressure is maintained 
by the pumps, the com- 
paratively small quan- 
tity of oil required for 
operating the valve 
mechanism passes to the 
relay cylinder, from 
which it exhausts to the 
cooler. In the larger 
machines an auxiliary oil 
pump is furnished for 
establishing a circulation 
with the turbine at rest. 

The glands on both 
the non-condensing and 
the condensing units are 
water sealed. This seal is 
effected by small bronze 
impellers fitted on either 
end of the turbine shaft 
and which revolve in an- 
nular chambers. Water 
is fed into these cham- 
bers and the centrifugal 
action of the impellers 
maintains a pressure 
which effectually seals 
the glands against air 
leakage into the casing 
or steam leakage into the 
atmosphere. The amount 
of sealing water required 
is very small. The 
grooves and mating col- 
lars on the balancing piston constitute a labyrinth packing. 

214. Allis-Chalmers Steam Turbine. — Fig. 270 shows a section 
through an Allis-Chalmers standard steam turbine, which is of the 




r* 

I 



474 STEAINI POWER PLANT ENGINEERING 

Parsons type but differs from the original Parsons machine and the 
Westinghouse-Parsons construction principally in manufacturing de- 
tails. In the older Parsons type, three balance pistons are placed at 
the high-pressure end. In the Allis-Chalmers design, the larger piston 
is placed at the low-pressure end of the rotor, behind the last row of 
blades, the other two remaining at the high-pressure end. This con- 
struction permits of a smaller balance piston and allows a smaller 
working clearance in the high-pressure and intermediate cylinders. 
In the Allis-Chalmers turbine the roots of the blades are dovetailed 
and fitted into a foundation ring, and the tips are incased in a channel- 
shaped shroud ring, thereby insuring a rigid and positively spaced 
construction. The governor is of the Parsons type, except that the 
main valve and pilot valve are actuated by hydrauKc instead of steam 
pressure. The bearings are of the self-adjusting ball-and-socket 
pattern and are kept ''floating in oil" by a small pump geared to the 
turbine shaft. The oil is passed through a tubular cooler with water 
circulation after it leaves the bearings and is used over and over again. 
215. Westinghouse Impulse-reaction Turbine. — With ths exception 
of a non-condensing unit and the purely impulse type described in 
paragraph 204 all high-pressure single-cylinder turbines constructed 
by the Westinghouse Company are of the combined impulse and re- 
action types. A typical unit is illustrated in Fig. 271. There are two 
rows of moving blades or buckets upon the impulse wheel with an in- 
termediate set of reversing blades, the operation being practically the 
same as in the first stage of the Curtis turbine. The drop in pressure 
in the nozzles is such that approximately 20 per cent of the total energy 
developed is absorbed by the impulse element. The steam discharged 
from the impulse element is expanded through the reaction elements 
in the usual manner. The substitution of the impulse element for the 
high-pressure section of reaction blading has no influence on the effi- 
ciency but results in a shorter machine and gives a more rigid design 
of rotor. From Fig. 271 it will be seen that the cylinder has been 
shortened not only by the substitution of the narrow impulse element 
for a comparatively wide section of reaction blading but also by the 
elimination of the intermediate balancing pistons or dummies as used 
in the conventional Parsons design. A further inspection of Fig. 
271 win show that the glands on each end of the cyhnder are subjected 
to exhaust pressure and that leakage of air into the turbine casing is 
prevented by the water-sealing device described in the preceding para- 
graph. Steam leakage past the balancing piston through the laby- 
rinth packing escapes into the exhaust. Fig. 272 shows a section 
through a double-flow impulse-reaction turbine which differs from the 



STEAM TURBINES 



475 




o 



fafl 

C! 
CO 



3 

a 

a 

g 
I 

O 

O 

c3 
O 



476 



STEAM POWER PLANT ENGINEERING 




Y}///rY//M///////M 



v W//''m :^. 



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'h/y////////.w/m//y/M 






STEAM TURBINES 477 

one illustrated in Fig. 271 by the use of reaction elements on either side 
of the impulse wheel. It will be seen from Fig. 272 that the double- 
flow is shorter than the single-flow machine by a length equivalent to 
that of the balance piston. Single-cylinder, double-flow impulse- 
reaction turbines have been constructed in sizes up to 22,000 kilowatts. 

216. Westlnghouse Compound Steam Turbines. — In a turbine maxi- 
mum centrifugal stresses occur at the exhaust end where the large volume 
of steam requires the greatest blade area. In the high-pressure blading 
involving the use of high-density steam the best velocity ratio con- 
ducive to high economy cannot be met by the rotative speed as de- 
termined by the exhaust end. To avoid a compromise with its resulting 
reduced efficiency the expansion is carried out in two or more separate 
elements. All of the modern large turbines built by the Westing- 
house Company are of the multi-cylinder type. The two-cylinder 
machines are arranged either tandem or cross compound. Three- 
cyHnder cross-compound units have also been built and it is not unlikely 
that the four-cylinder construction may be used for very large units. 
The advantages of the multi-cylinder construction over a single cyhnder 
of like capacity are as follows: 

1. Smaller cylinder structure. 

2. Lower temperature range within the cylinder. 

3. Highest efficiency for each cyhnder for the expansion of range 
involved. 

4. Reduction in weight of the parts to be handled. 

5. Possibility, in case of emergency, to operate either cylinder alone. 

In the Westinghouse compound turbine the high-pressure element 
is practically a typical single-cylinder reaction turbine and the low- 
pressure element is a reaction turbine of the double-flow type. The 
high-pressure element of the 30,000-kw. turbine at the 74th Street 
Station of the Interborough Rapid Transit Company, N. Y., operates 
at 1500 r.p.m. and the low-pressure element at 750 r.p.m. 

217. Elementary Theory. — Reaction Turbine. — Fig. 273 gives a diagram- 
matic arrangement of fixed and stationary blades in the first stage of 
a multi-stage reaction turbine. The steam enters the stationary blades 
at a comparatively low initial velocity and is there partially expanded 
and impinges against the moving blades at velocity Vi. In practice 
Vi is made such that there is practically no impulse when the jet strikes 
the vanes. In passing through the moving vanes the steam is further 
expanded and leaves at absolute velocity V-z, exerting a reactive force 
on the rotor. The steam enters the second set of stationary blades 
with absolute velocity V2 and is still further expanded to velocity V3, 
and so on. 



478 STEAM POWER PLANT ENGINEERING 

The energy imparted to the steam in the first set of stationary blades 



IS 



W W 



(187) 



in which 
//i = initial heat content, B.t.u. per lb., 

H2 = heat content after expansion through the blades, B.t.u. per lb., 
W = weight of steam, lb. per sec, 
Vi = velocity imparted to the jet by expansion. 

The absolute spouting velocity Vs = Vq -\~ Vi, in which Vq = en- 
trance velocity to the fixed blades. 




Fig. 273. Velocity Diagram, Westinghouse Reaction Turbine. 
The energy imparted to the steam in the first set of moving blades is 



W 



(188) 



E2 = ^^W-v,'), 
in which 

Vi = relative velocity of steam entering the moving blades, 
V2 = relative velocity of steam leaving the moving blades. 

The total energy available in the first stage is Eg + Eo, in which 
Eg = kinetic energy of the jet leaving the stationary vanesi Es =9" ^5^ )• 
The energy converted into useful work in this stage is 



W 



E = E,-\-E2-^V2' 
= iVs' + V2' - V,' - V2') 



W 



(189) 



V2 = absolute velocity of the steam leaving the moving blades. 
This residual velocity will also be the initial entrance velocity of the 
second stage. 

Each stage may be analyzed in a similar manner. 



STEAM TURBINES 479 

Example 33. Construct the velocity diagram and calculate the 
work done per stage in a frictionless reaction turbine for the following 
conditions: Heat drop per stage = 18 B.t.u. per lb. of steam; periph- 
eral velocity = 300 ft. per second; exit angle = 30 deg.; entrance 
velocity = zero. 

The velocity imparted to the steam in the first set of stationary 
blades is 

Vi = 224 y = 672 ft. per sec. 

The spouting velocity is 

V, = Vi = 672 ft. per sec. 

Lay off Vs in direction and amount and combine with u = 300, 
Fig. 273. The resultant is Vi, the velocity of the steam relative to the 
blades. The angle between Vi and the hne of motion of the wheel 
will be the entrance blade angle. From the diagram Vi = 438. The 
energy given up by expansion in the moving blades is 

1 o 

£- = 778 X -^ = 7002 ft. per sec. 
Substituting ^i = 438 and E^ = 7002 in equation (188) 

7002 =^U^- 438^), 
V2 = 802 ft. per sec. 

The resultant of V2 and u is V2, the residual velocity of the steam 
leaving the moving blades. From the diagram Vo = 576. 

The energy converted into work in the first stage is from equation 

(189) 

E = (672% 8O2' - 438' - 576') ^ 

64.4 

= 10,420 ft. lb. per sec. for each lb. of steam flowing 
through the turbine. 

In the actual turbine the various friction and leakage losses must 
be included in the calculation. Such an analysis is beyond the scope 
of this text and the reader is referred to the accompanying bibliography. 

218. Exhaust-steam Turbines. — Low and Mixed Pressures. — In a 

general sense the reciprocating engine reaches its maximum overall 
economy at a vacuum of about 26 inches referred to a 30-inch barom- 
eter. Cylinder condensation and excessive size of low-pressure cylin- 
der usually offset the reduction in steam consumption for vacua higher 
than 26 inches. When it is considered that an increase in vacuum 
from 26 to 29 inches (initial pressure 200 lb. abs.) increases the avail- 
able energy about 22 per cent the loss in economy due to the inability 
of the engine to utilize the higher vacuum is at once apparent. The 



480 STEAM POWER PLANT ENGINEERING 

ability of the turbine to take care of large volumes of steam and to 
avoid cylinder condensation makes a high vacuum desirable for eco- 
nomical reasons. The gain possible by taking advantage of high 
vacua has brought about the exhaust-steam turbine. Since the in- 
stallation of the low-pressure turbine connected to the 7500-kw. angle- 
compound engine at the 59th Street Station of the Interborough Rapid 
Transit Company of New York, exhaust-steam or low-pressure tur- 
bines have been installed in many plants. In this noteworthy instal- 
lation the addition of the low-pressure turbine effected : 

a. An increase of 100 per cent in maximum capacity of plant. 

h. An increase of 146 per cent in economic capacity of plant. 

c. A saving of approximately 85 per cent of the condensed steam 
for return to the boiler. 

d. An average improvement in economy of 13 per cent over the best 
high-pressure turbine results guaranteed at that time. 

e. An average improvement in economy of 25 per cent over results 
obtained by the engine units only. 

/. An average unit thermal efficiency of 20.6 per cent between the 
limits of 6500 kw. and 15,500 kw. 

The low-pressure turbine is installed between the exhaust of the 
low-pressure engine cylinder and the condenser as shown in Fig. 274. 
Running with the engine the low-pressure turbine generator carries a 
variable load without governor control. The turbine generator takes 
care of the speed by automatically taking such a load as will keep the 
frequency in unison with that of the engine-driven unit. The turbine 
is equipped with the usual emergency speed-limit attachment for 
cutting off the steam supply should the speed exceed a predetermined 
limit. 

Although numerous examples may be cited showing great gains in 
both capacity and economy in existing reciprocating-engine plants 
by the addition of a low-pressure turbine, a combined reciprocating- 
engine low-pressure turbine unit would not be selected in place of a 
high-pressure turbine unit for a new plant. Combined units of large 
power will cost approximately $40 per kw. against $8 per kw. for 
the high-pressure turbine unit. The space requirements of the com- 
bined unit are much larger than that of the high-pressure turbine unit 
and the cost of attendance, supphes and maintenance is also greater. 
Besides, high-pressure turbine economy has been greatly improved 
and water rates considerably below that guaranteed at the time the 
low-pressure turbines were installed in the 59th Street Station are 
realized in current practice. There is no question as to the economy 
eft'ected in adding an exhaust steam turbine to large non-condensing 



STEAM TURBINES 



481 




Fig. 274. Low-pressure Turbine Installation at the 59th Street Station of the Inter- 
borough Rapid Transit Company, New York. 




Fig. 275. Westinghouse Double-flow Low-pressure Turbine. 



482 



STEAM POWER PLANT ENGINEERING 



reciprocating engines, but with condensing engines it should be borne 
in mind that without increase in vacuum the addition of a low-pressure 
turbine will hardly warrant the extra expense. 

Exhaust steam turbines may be divided into three classes, the 
division depending on the supply of low-pressure steam, viz., straight 




Fig. 276. Rateau Low-pressure Steam Turbine Installation. 



low-pressure, mixed-pressure, and high-and-low-pressure turbines. Low- 
pressure turbines use exhaust steam only and are installed where 
there is an ample supply of low-pressure steam to carry the load at 
all times. Mixed-pressure turbines carry the full load (1) on all low- 
pressure steam, (2) all high-pressure steam, (3) any proportion of 
high-and-low-pressure steam at the same time. High-and-low-pressure 
turbines can carry the load on either high-pressure or low-pressure 
steam, but are not arranged to carry the load on both high- and low- 
pressure steam at the same time. 

Low-pressure turbines may be installed so as to receive exhaust 
steam from a number of engines and other steam-actuated appliances, 
all of which exhaust into a common main or receiver, or they may 
be installed so as to receive the exhaust from one engine only. 

Low-pressure turbines are frequently installed in connection with 
regenerator accumulators, to rolling-mill engines, steam hammers, and 
other appliances using steam intermittently, and have proved to be 
paying investments. The generator accumulator is intended to regu- 
late the intermittent flow of steam before it passes to the turbine. 
The steam collects and is condensed as it enters the apparatus and is 
again vaporized during the time when the exhaust of the engines dimin- 
ishes or ceases. 



STEAM TURBINES 



483 



The regenerator usually consists of a cylindrical boiler-steel shell 
divided into two similar chambers by a central horizontal diaphragm, 
Fig. 277. In each compartment arc a number of eUiptical tubes A, 
each of which is perforated with a number of f-inch holes. The spaces 
surrounding the tubes and, under certain conditions, the tubes them- 
selves are filled with water to a height of about four inches above the 
top of the upper tubes. Baffle plate B serves to separate the entrained 
moisture from the steam. The operation is as follows: Exhaust steam 
enters the apparatus at N, passes to the interior of the elliptical tubes, 
and escapes into the steam space through the perforations and thence 



Exhaust Steam Inlet 




Fig. 277. Rateau Regenerator Accumulator. 



to the turbine. When the supply of steam from the main engine 
ceases, the pressure in the regenerator decreases, the water Hberates 
part of the heat it has absorbed and a uniform flow of low-pressure 
steam is given off. The continued demand of the turbine reduces the 
pressure in the accumulator and causes the steam still retained in the 
tubes to escape, thereby maintaining the circulation of the water 
(indicated by arrowheads) and facilitating the liberation of steam. 
Suitable valves regulate the limits of pressure in the accumulator and 
prevent the return of water to the main engine. 

In the size normally installed this type of accumulator will furnish 
a sufficient supply of steam for four minutes with exhaust entirely 
cut off. If the period is longer than four minutes it becomes necessary 
to admit live steam. Low-pressure turl^ines develop one electrical 
horsepower-hour on a steam consumption of about 30 pounds with 
initial pressure of 15 pounds absolute and a back pressure of 1.5 pounds 
absolute. Fig. 278 gives the perforaiance of a typical Westinghouse 



484 



STEAM POWER PLANT ENGINEERING 



low-pressure turbine for various vacua, initial pressure 15 pounds 
absolute. 

The weight of water W required to operate the low-pressure turbine 
for a given period with a predetermined temperature drop may be 
calculated from the relationship 

W = -^'^, (190) 

Qi - q2 

in which 

t = maximum number of minutes the exhaust supply may be en- 
tirely cut off, 
s = water rate of the turbine, pounds per minute, 
r = mean latent heat at regenerator pressure, 
Qi = heat of the liquid corresponding to maximum temperature of 

water in regenerator, deg. fahr., 
g2 = heat of the Hquid corresponding to minimum temperature of 
water in regenerator, deg. fahr. 



75 




1 




Economy Test 
1500 B.H.P. Westinghouse Low Pressure Turbine 








Water Kates per B.H.P. Hr. at different Vacua 
Speed 1800 R.P.M. 


70 


























w 


























65 

W60 

W55 

30 
35 




V 














0^ *^ 


Avft^ 


'^^ 


f^^ 


" 






\\ 


\ 






V^ 




iii-" 














V 


A 
























\ 


V 


\ 


2 


i" 


r 


26" 
FulijLoad 


2 


7" 


2 


8" 






V 


\ 






V 


acuun 


L in in 


.•lies(t 


„ 

Bai 


0.) 










\ 


N 


V 


-^. 


"'• fi. 






















\ 






•=i^ 


^ 


^£2C 


£.«., 


L=^i- 
















^^ 








->- 




-Jil 


r— 


= 

























^C 




20 





































100^ 
90^ 
80^ 



200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 
Load - Brake Horse power 

Fig. 278. Performance of Westinghouse Low-pressure Turbine. 

If the regenerator is to absorb M pounds of exhaust steam in t minutes 
as in case of a sudden flux of exhaust the weight of water Wx required is 

Mr 
Wr=-^^^^^- (191) 

Example 34. Determine the weight of water to be stored in a re- 
generator to operate a 500-horsepower exhaust steam turbine for five 



STEAM TURBINES 



485 



minutes if the steam supply is entirely cut off; pressure drop 17 to 14 
pounds absolute, turbine water rate 30 pounds per horsepower-hour. 

^ , 500 X 30 „_ 965.6 + 971.9 „_ ^ 
^ = 5, s = — = 2o0, r = 2 " = ^^^-^f 



qi = 187.5, ^2 = 177.5, 
5 X 250 X 968.8 



W 



187.5 - 177.5 



121,100. 



If the regenerator is to absorb 2000 pounds of the exhaust steam in five 
minutes during a period of sudden flux, 

_ 2000 X 968.8 iQ^7.n 
^^^ ^ 187.5 - 177.5 = ^^^-^^Q- 

Theory of Steam Accumulators and Regenerative Processes: F. G. Gasche, Proc. 
Eng. Soc. Wes. Penn., Dec, 1912, p. 723. 




Fig. 279. Westinghouse Mixed-pressure Turbine. 

In the mixed-pressure turbine the transition from all low pressure 
to all high pressure, through all the conditions intermediate between 
these extremes, is provided for automatically by the turbine governor; 
a deficiency of low-pressure steam causes the high-pressure nozzles 
to open automatically. With this arrangement it is not necessary 
for purposes of economy to proportion exactly the low-pressure turbine 
to the amount of exhaust steam available, but within limits it may 
be made as large as the load demands. 

Mixed-pressure turbines have been constructed in single units as 
large as 10,000 kw. 

The high- and low-pressure turbine is used when there is a suflficient 
supply of low-pressure steam to carry the load for a long period, say 
three or four months, and when for a similar period only high-pressure 
steam is available. When designed for this pressure range the tur- 



486 



STEAM POWER PLANT ENGINEERING 



bine does not operate at maximum efficiency at either the high- or the 
low-pressure condition. For this reason it is doubtful whether this 
arrangement results in better overall economy than two separate units, 
a high- and a low-pressure turbine. 

319. Advantages of the Steam Turbine. — The principal advantages 
of the steam turbine are: (1) low first cost; (2) low maintenance and 
attendance; (3) economy of space and foundation; (4) absence of oil 
in condensed steam; (5) freedom from vibration; (6) uniform angular 




Fig. 280. Section through a Westinghouse Bleeder Type Turbine. 



velocity, and (7) high efficiencies for large variations in load. The 
reciprocating engine is well adapted for pumping stations, compressor 
plants, hoisting engines, and the hke, requiring low angular velocity, 
and for reversing service, but its place is being rapidly taken by the 
steam turbine for alternating-current dynamos, centrifugal pumps and 
blowers requiring high angular velocity. The recent development of 
high-efficiency speed-reduction gearing makes it possible for the tur- 
bines to compete with the engines for low-speed work. In fact the 
geared turbine is rapidly replacing the engine for low-speed work. 

First Cost. — Because of the various uses to which turbines are ap- 
plied and on account of the extreme variation in design general rules 



STEAM TURBINES 487 

for approximating^ the cost of turbines are without purpose. Values 
based on rated capacity vary within such wide hmits that average 
figures are apt to lead to serious error. In a general sense steam tur- 
bines are lower in first cost than steam engines of equivalent rated ca- 
pacity irrespective of size. Specific figures are given in Chapter XVIII. 

Maintenance and Attendance. — Although composed of a large num- 
ber of parts as compared with a reciprocating engine of the same capacity, 
there are few moving parts and rubbing surfaces. The only contact 
between rotor and stator is in the main bearings, and the problem of 
lubrication is therefore a simple one. The absence of pistons, stuffing 
boxes, dish pots, etc., reduces the cost of maintenance and attendance 
to a minimum and limits the possibility of leakage. See Chapter XVIII 
for specific figures. 

Economy of Space and Foundation. — The floor space required by 
practically all types of turbines is considerably less than the space 
requirements of piston engines. Vertical three-cyUnder compound 
Corliss engines of the New York Edison type require the least floor 
space of any large slow-speed reciprocating engines, but take up about 
twice the space of a Parsons turbine installation of the same size. With 
non-condensing high-speed engines the comparative economy in space 
is less marked. The average space occupied by turbine units is approxi- 
mately I less than that of engine units of equivalent capacity, but specific 
cases may be cited in which the ratio varies widely from the average. 
In the modern central station the actual space reduction per kilowatt 
of plant rating is much less than that referred to the prime mover 
only because of the tendency toward less crowded conditions. 

The weight of the steam turbine is very small compared with a re- 
ciprocating engine of the same horsepower. The New York Edison 
engine and generators weigh more than eight times as much as a turbine 
installation of equal capacity. The turbine, for this reason, and also 
because of the total absence of vibration, requires a relatively light 
foundation. In many instances the foundation consists of steel beams 
with concrete arches sprung between them resting upon the floor, and 
the basement underneath may be used for the condenser instead of the 
massive foundation required for the reciprocating engine. Engines are 
seldom constructed in sizes above 5000 horsepower, whereas single 
turbine units of 30,000 kw. are not uncommon and a turbine 60,000 
kw. normal capacity is now being installed in the 74th Street Station 
of the Interborough Rapid Transit Co., N. Y. 

Absence of Oil in Condensed Steam. — As the steam turbine requires 
no internal lubrication, oil does not come in contact with the steam, 
and the condensed steam from the surface condensers is available for 



( 



488 STEAM POWER PLANT ENGINEERING 

boiler-feeding purposes without purification. In many cases the re- 
use of condensed steam effects a large saving in cost of feed water and 
in expense for maintenance and cleaning of boilers. The amount of 
entrained air is reduced to a minimum and consequently the work of 
air pumps is lessened. 

Regulation. — The variable pressure at the crank pin of a recipro- 
cating engine necessitates the use of a heavy flywheel to keep the in- 
stantaneous angular fluctuation within practical limits. In the steam 
turbine the motion is purely rotary and a flywheel is not necessary. 
In the former there are always instantaneous variations in velocity 
during each revolution, even with constant load, while in the latter the 
speed is practically constant. A number of published tests of Par- 
sons and Curtis turbines show an average fluctuation of 2 per cent from 
no load to full load and 3 per cent from no load to 100-per-cent over- 
load. Although closer regulation than this is possible, it is not deemed 
necessary, particularly in alternating-current work where a compara- 
tively wide range is desirable for parallel operation. 

Overload Capacity. — The overload capacity of any prime mover de- 
pends entirely upon the designation of the rated load. The maximum 
economy of the average piston engine lies between 0.7 and full load, 
and for this reason the rated load refers usually to this maximum eco- 
nomical load. Evidently if the engine is rated under its maximum 
possible output it is capable of overload. Under the existing system of 
rating the average piston engine is capable of operating with overloads of 
25 to 50 per cent. According to the old rating the steam turbine was 
capable of overloads ranging from 100 to 200 per cent and much con- 
fusion arose in determining the station load factor. Current turbine 
practice gives as the normal rating the maximum continuous load which 
can be carried for 24 hours when under control of the primary valves. 
Through the agency of the secondary valves overloads of 50 per cent or 
more are possible. The steam economy of the turbine is superior to 
that of the engine for overloads. Since all modern turbines are designed 
for a point of best steam consumption somewhere regardless of what 
their rating may be, the actual rating means little. 

230. Efficiency and Economy of Steam Turbines. — A general com- 
parison of the water rates of piston engines and steam turbines is very 
unsatisfactory because of the wide range in operating conditions. In 
a general sense the piston engine is more economical in the use of steam 
than the turbine for non-condensing service and the reverse is true for 
high-pressure, high-vacuum condensing service. Condensing engines 
of the uniflow or poppet-valve type have shown superior economy 
(under favorable conditions) to the turbine for sizes up to 3000 horse- 



STEAM TURBINES 



489 



power and in some instance's up to 5000 horsepower but heat economy 
is only one of the many factors entering into the ultimate cost of power. 
For sizes over 3000 horsepower the turbine is in a class of its own and 
piston engines above this size are seldom found in modern stationary 
practice. A comparison of the curves in Fig. 215, showing typical 
economy curves of high-speed single-valve non-condensing engines, and 
of Fig. 283, showing the performance of non-condensing steam turbines, 
is somewhat in favor of the piston engine, the difference decreasing as the 
size of unit increases. A similar comparison of the performance curves 



t 1 






























1 


1 
















V 


5h. 


p.- 


^31 


>. 














































\ 
























TV 


^ate 


rE 


ate 


in 


LbJ per Brak 
-Full- Load - 
essure 160 Hb. ( 


eH 


.p. 


rH 


r. 








50 


b.p. 


-38 


lb. 




















In 


itia 


Pr 


Jauge 












] 


1 


























St 
Br 


earn Dry | 
ck Pressure 


I 
L Lb. G 


auge 














V 
























































\ 


100 


h.p 


,-3 


21b. 


















































\ 


























































^ 


^ 


)h.i 


).- ; 


91b 
























































-^ 


^ 





500 


h.p 


-2 


rib. 




_75C 


















































*~~ 


— ' 






___^ 


loop h. 


). -5 


J4.75 lb. 












































" 








: 


500 


M 


^ 


lb. 


















NC 


N- 


CO 


VID 


EN 


SIN 


G 


5TE 


lAM 


T 


JR 


BIN 


ES 


























































































_ 







































32 



I 

es 
^20 



200 400 600 800 1000 1200 

Eated Full Load -Brake Horse Power 

Corrections for fractional loads. — Increase full load water rate as follows: ^ — 
20%; f-8%; 1-0%; U — 5%. 

Corrections for initial pressures. — 175 lb. deduct 3%; 200 lb. deduct 5%; 125 lb. 
add 5%; 100 lb. add 10%; 75 lb. add 20%. 

Corrections for increased back pressure. — Add for each lb. back pressure 200 lb. 
— 1%; 175 1b. — U%; 150 lb. — U%; 125 1b. — 2%; 100 1b.— 2A%; 75 1b.— 
3%. 

Correction for superheat. — Subtract 1% for each ten degrees superheat up to 200 
degrees. 

Fig. 281. Average Water Rates of High-grade Small Non-condensing Steam 

Turbines. 



of compound single-valve, single-cyHnder four-valve, and compound 
four-valve non-condensing piston engines with those of steam turbines 
of the same size show marked increase in economy in favor of the piston 
engine. For sizes between 2000 and 6000 horsepower there is little 
difference between the steam economy of the very best grade of piston 
engine and that of the turbine. Piston engines above 10,000 horse- 
power have not been built for stationary practice, hence a comparison 
with the turbine for larger sizes is impossible. The Manhattan type at 
the 74th Street Station of the Interborough Rapid Transit Company 
represents the largest piston engines (7500 kw.) ever constructed for 



490 



STEAM POWER PLANT ENGINEERING 



central station service. The heat consumption of these engines is 
considerably more than that of the modern turbo-generator of the 



12.0 

a 
o ^ 

a I 

a W 

o I, 

o 5 

£ ^ 
|;on.o 















































































^-^ 


^^ 








































' 






"--- 






■ 




' 







^ 


























♦ 




Water Rate Corrected to 
215 lb. per sq. in. Abs. 
120 Deg. Fahr. Superheat 
2"-In. Vacuum 
























































1 





16 



22 24 26 28 

Load in Thousands of Kilowatts 



Fig. 282. 



Performance of 30,000-kw. Westinghouse Compound Turbine, Inter- 
borough Rapid Transit Co. 



same capacity. Tables 81 and 85 give the general conditions of op- 
eration and the steam consumption of exceptionally good piston engines 
of various sizes and types and Table 89 similar data of first-class tur- 



TABLE 89. 

PERFORMANCE OF THE MODERN STEAM TURBINE AT RATED CAPACITY. 
(Manufacturer's Guarantee.) 



Index. 


Make of Turbine. 


Rated 
Capacity. 


R.P.M. 


Initial 
Pres- 
sure, 
Lb. 
Abs. 


Back 
Pres- 
sure, 
Inches. 


Super- 
heat, 
Deg. 
Fahr. 


Lb. Steam 

Per 

Kw-hr.* 


Rankine 
Cycle, 
Ratio. 


1 


Westinghouse 


500 Kw. 


3600 


165 


2.0 





19.8 


t53.0 


2 


'' 


1,000 " 


3600 


165 


2.0 





18.1 


t58.0 


3 


<< 


5,000 " 


1800 


165 


2.0 





16.0 


t65,6 


4 


(( 


15,000 " 


1800 


215 


1.5 


125 


12.6 


t71.0 


5 


" 


30,000 " 


1200 


235 


1.0 


200 


10.65 


t75.0 


6 


" 


45,000 " 


1200 


215 


1.0 


200 


10.65 


t76 


7 


Curtis 


500 " 


3600 


215 


2 





18.5 


t54.0 


8 


" 


1,000 " 


3600 


215 


2.0 





17.5 


t57.1 


9 


" 


5,000 " 


1800 


215 


2.0 


125 


14.3 


t64.6 


10 


" 


15,000 " 


1800 


215 


2.0 


125 


12.5 


t74.0 


11 


u 


30,000 " 


1200 


215 


2.0 


125 


12.2 


t75.8 


12 


iC 


45,000 " 


1200 


215 


2.0 


125 


11.9 


t77.6 


13 


Kerr 


25 Hp. 


3600 


165 


At. 





t43 


131 


14 


i( 


50 " 


3600 


165 


At. 





t38.0 


t34.2 


15 


" 


100 " 


3600 


165 


At. 





t32 


J40.6 


16 


'' 


200 " 


3600 


165 


At. 





t29.0 


$44.9 


17 


(< 


500 " 


3600 


165 


At. 





t27.0 


t48.1 


18 


(I 


750 " 


3600 


165 


At. 





125.5 


t51.0 


19 


" 


1,000 " 


3600 


165 


At. 





$24.75 


157.5 


20 




1,500 " 


3600 


165 


At. 





124.0 


159.1 



* Lower water rates than these have been guaranteed for higher pressures and superheat and for lower 
back pressures. The guaranteed water rate for a 20,000-kw. Curtis turbo-generator at the River Plant 
of the Buffalo General Electric Company is 10.6 lb. per kw-hr. for initial pressure, 265 lb. abs., back pres- 
sure 1 in., superheat 250 deg. fahr. 

t Based on electrical horsepower. 

:{: Based on developed horsepower, 



STEAM TURBINES 



491 



bines. A study of these tables will show that the choice must be based 
on other factors than the steam consumption. In a general sense the 
piston engine is superior to the turl)ine for high back pressures, slow 
rotative speeds, reversing service and heavy starting torques, while 
the turbine has practically superseded the piston engine for large cen- 
tral station units and for auxiharies requiring high rotative speed. 
Recent tests of geared turbines show exceptionally high efficiency for 
sizes as large as 10,000 hp., and it is not unUkely that the turbine 
equipped with this device will offset the low rotative speed factor of 
the piston engine. 

If the tests of steam turbines and piston engines could be made at 
some standard initial pressure, back pressure and quaUty or superheat, 
then a comparison could readily be made, but both types of prime 
movers are designed to give the best results for special operating con- 
ditions, and any marked departure from these conditions will result 
in loss of economy. It is frequently desired, however, to make a 
comparison between the economy of the different machines, and the 
following methods are in vogue: 

(1) Steam consumption under assumed conditions. 

(2) Heat consumption per unit output per minute above the ideal 
feed-water temperature. 

(3) Rankine cj^cle ratio. 

Steam Consumption under Assumed Conditions {Standard Correction 
Curves) : This method for comparing engines or turbines or both is 
best illustrated by a specific example: 

Example 35. Compare the full-load performance of a 125-kilowatt 
direct-connected piston engine with that of a 125-kilowatt turbo- 
generator with operating conditions as follows: 



Engine.. 
Turbine 



Steam Consump- 
tion, Lb. per 
Kw-hour. 



25.0 

22.7 



Initial Pres 
sure, Lb. 
Absolute. 



160 
110 



Vacuum, 
Inches of Hg. 



25.5 

28.0 



Superheat, 
Des. Fahr. 




125 



Manufacturers of steam turbines have provided correction curves as 
illustrated in Fig. 284, showing the influence of varying vacuum, super- 
heat and pressures on the steam consumption.* From curve B, we 
find that the steam consumption of the turbine should be decreased 
2.5 pounds to give the equivalent at 160 pounds initial pressure; from 
curve A it should be increased 2.5 pounds to give the equivalent at 



* These curves are drawn to a much larger scale than the reproduction given here. 



492 STEAM POWER PLANT ENGINEERING 

25.5 inches of vacuum, and from curve C it should be increased 2.5 
pounds to give the equivalent at degree superheat. The full-load 
steam consumption for the turbine under the engine conditions is there- 
fore 22.7 - 2.5 + 2.5 -f- 2.5 = 25.2 pounds per kilowatt-hour. 

The ratio method is also used in this connection, thus: The full-load 
steam consumption at 160 pounds pressure, curve B, Fig. 284, is multi- 

25 
plied by the ratio ^^^-^ to give the equivalent consumption at 110 pounds 

(25 is the steam consumption at 160 pounds and 27.5 the consumption 

at 110 pounds). Similarly, the correction ratio to change the con- 

25.5 
sumption at 28 inches of vacuum to 25.5 is -^ , and to correct 125 

25 
deg. fahr. superheat to deg. fahr. is ^-^- 



Summary. 

25 
Pressure correction ^r^^ = 0.91 = — 9 per cent. 

27.5 
Vacuum correction -^^ = 1.10 = 10 per cent. 

25 

Superheat correction ^r^ = 1.11 = 11 per cent. 

Net correction 12 per cent. 

Corrected steam consumption = 22.7 + 0.12 x 22.7 = 25.4 pounds per 
kilowatt-hour. 

The ratio method is generally used if the difference between the 
corrected steam consumption and that of the correction curves for 
the same conditions is greater than 5 per cent (''The Steam Turbine," 
Moyer, p. 128). 

This ratio method for correcting steam consumption at full load may 
be used without appreciable error for half to one and one-half load and 
is the only practical method for quarter load. (Engineering, London, 
March 2, 1906.) 

Heat Consumption: 

The heat consumption B.t.u. per unit output per minute above the 
ideal feed-water temperature may be expressed 





W {H, ■ 


- qi) 






60 






For the case 


cited above 






Engine, 


25 (1194.1 - 
60 


98) 


455 B.t.u. 


Turbine, 


22.7 (1264.2 - 


70) 


451 B.t.u. 



STEAM TURBINES 



493 





- 


"■ 






~ 




~ 


^ 




T 


~ 




~ 






"■ 




~ 




~ 










~ 








"" 




~ 




~ 








■" 


— 1 




~ 


-^ 


140,000 
















































































<^^ 










©^-Guarantee points 
A - Water rate-low vacuum as tested 
B'- Water rate-high vacuum as tested 
C - Water rat«- corrected to 175 lb. pressure, 
100 deg. superheat, 28 in. vacuum 


































.^ 








130,000 




























^ 






r'M' 
































^ 






X 


/■ 
















120.000 






















»^ 






^. 


^ 






























A- . 


/^ 








<^ 


^ 


6 




















110,000 






















































y 






XI 


■> 










































































y 


B- 






•^ 




























u 
















































^^ 




1 


5-: 


r^ 


1 












































































y^ 








[^ 




r 
































S. 90.000 


















































y 




















































































^ 


pi 


' 








































3 80,000 












































f 


Lx 












































^ 






































^ 


v" 
















































1 70,000 






































,-» 


' 




















































































•' 








n 














































Total Vi 






























^ 


y 
















^ 










/ 


























































,-' 


























^ 


r^ 


k 






















































, 
















c 


) 


























































































^ 


^ 
























____ 






































,^ 
























"^ 


^ 




^ 
^ 








B\ 
















































' 
































■^ 






^ 








— 




=« 


^ 


'j 




















r^nr, 












^^ 










































r 


p/ 


r— 


->. 












® 
















,* 






























































" 








~ 






__ 




- 


20,000 




>' 










































































































































































10,000 






















































































































































































_ 












- 
















L_ 



























































18 
17 5 



1,000 2,000 3,000 4,000 6,000 6,000 7,000 8,000 9,000 10,000 11,000 

J.oadJn Kw, 

Fig. 283. Performance of lO.OQO-kilowatt Westinghouse Double-flow Turbine, 
City Electric Co., San Francisco, Cal. 



100 



no 



steam Pressure, Lbs. per Sq. In. Absolute 
120 130 140 150 160 170 



180 



190 



200 



S34 



32 



«28 
§26 

|24 
§22 



g20 



A OF Superb eat, 165^ Abs, 

B •' .. , 28 in Vacuum 

C 165* Absolute, " " 




TYPICAL CORRECTION CURVES 

FOR 125 K.W. STEAM TURBINE 

FULL LOAD CONDITIONS. 



20 



40 


CO 


80 100 120 140 
buperheat, Deg. Fab. 


100 


180 


200 


21 


22 


23 24 25 26 
Vuciauin, lucbes of Mercury 


27 


28 


29 



Fig. 284. 



494 



STEAM POWER PLANT ENGINEERING 



Rankine Cycle Ratio: 

The Rankine cycle ratio, or the extent to which the theoretical possi- 
bihties are realized, may be expressed 



Er = 



For the case cited above 
Engine, 

Turbine, 



2546 X 1.34 
W {H, - H2) ' 

2546 X 1.34 
25(1194.1 - 915) 
2546 X 1.34 
22.7 (1264.2 - 915.3) 



0.49. 



0.43. 



In the assumed case the turbine is the more economical in heat con- 
sumption, but the engine is the more perfect of the two as far as theo- 
retical possibilities are concerned. 

231. Influence of Superheat. — Theoretically, the gain in heat economy 
due to the use of superheat is the same for all prime movers of whatever 
type. In the actual mechanism the rate of improvement with increase 



Load in Thousands of Kilowatts -for Vacuum Correction only. 
12 14 16 18 20 22 • 24 26 28 



30 



-5 






\ 
























i 












y 
O 

1 

a 








\ 


















Standard Conditions 
Initial Pressure 215 Lb. per Sq. In. 
Superheat 120 Deg. Fahr. 
Vacuum 29.0 Inches 












\ 


















--. 


^ 


_Va 


2ua 


^ 


^ 


i-for 


each 


^O.-1-i 






















1.5 1 
-1.0^ 


c 

c-1 

o 

1 " 












■ , 



























-..5| 

0.0 g 
+ 0.5 3 


















"^ 


^~r^ 


^u, 


eCo 




































^K 




"T" 


^ 


■ — ■ 






















tion 








rs( 


,.„ Lis over .29.0 


u. 






— ■ — 




-=- 


1+3 

+5 
+6 


forP 


ach( 


).l i" 


. when V" 


^^'^ ' 


















Vnf 


^ 


orrec 














p^v 










































^ 












































•s 


N 












z 


































\ 










04 


















































5 

1€ 





7 
17 






9 
1^ 


U 

s 


Abs 


11 

upe 

19 

olut 




•Ilea 

ePr 


13 
t, D( 

20 
essu 




3gre 

re— 


IS 
esF 

21 
Lb. 




ahre 

per 


17 

nlie 
22 
Sq. 



it 

[n. 


If 






21 
24 






23 
25 









Fig. 285. Correction Factors for 30,000-kilowatt Westinghouse Compound 

Turbines. 



of superheat differs with type and size. The gain in the reciprocating 
engine is due mainly to the reduction of cylinder condensation while 
in the turbine the improvement in economy is due primarily to the 
reduction in "windage" and other friction losses. In both types the 
actual gain is much greater than in the perfect mechanism for pure 



STEAM TURBINES 



495 



adiabatic expansion. In the ideal or frictionless high-pressure condens- 
ing turbine an increase of 35 deg. fahr. superheat effects an increase of 
about 1 per cent in thermal efficiency. In the actual turbine the steam 
consumption is improved 1 per cent for every 6 to 14 deg. fahr. super- 
heat. The advantage of superheat is greater with non-condensing than 
with condensing units and is even more marked with low-pressure units. 
In average practice the maximum temperature of the steam is ap- 
proximately 500 deg. fahr., but in large central stations temperatures 




Full Load Operation 
175 llis.Steani Press 
-210-Peed-W-|iter-T4mp- 
Cp VJalues from | 

Knobjauch & Jakob 

Investment & Maintenance 
1 I Probated Adsumin'g 



qual-|Capital-&-F-u6l-Expfen8e— 



Note: 

No deductions made for 

increasing Radiation Idssea 



with inci eased superheat 



80 100 120 140 
Superheat-Deg".Fahr. 



220 



Fig. 286. Influence of Superheat on Overall Economy of Operation. 

of 600 degs. are not uncommon and a number of recent installations are 
designed for total steam temperature of 700 deg. fahr. The General 
Electric Company is prepared to design Curtis turbines for temperatures 
as high as 800 deg. fahr. should the demand warrant such high tem- 
peratures. The higher the initial temperature the greater will be the 
investment cost and a point is eventually reached where the increased 
fixed charges will offset the gain in heat economy. This is illustrated 
in Fig. 286, the curves of which though based on a specific case are 
appHcable in general principle to all cases. The influence of superheat 
on the economy of a 30,000-kw. turbo-generator is illustrated in Fig. 285, 



496 



STEAM POWER PLANT ENGINEERING 



323. Influence of High Initial Pressure. — The theoretical gain due 
to increase in initial pressure has been discussed in paragraph 179. 
In view of the marked improvement in heat economy for very high 
initial pressures it is only a matter of time when pressures far above the 
present maximum will be a matter of everyday practice. It seems 
that the only difficulty in the way of very high pressures is the boiler. 
There is no particular mechanical obstacle in designing the turbine 
since it simply means the use of heavier parts for the extremely high 
pressure end and a somewhat increased cost of construction. Because 

of the increased density of high- 
pressure steam the windage and 
friction losses would be correspond- 
ingly increased in the high-pressure 
stages and the dew point with its 
attending losses would be advanced 
farther up in the turbine. Just 
what will be the economical limit 
cannot be predicted with any de- 
gree of certainty. In case of ma- 
chines of 20,000-kw. capacity or 
more intended for operation in 
large power houses when the units 
are running at or near their most 
efficient load, considerable invest- 
ment is warranted for the sake of 
a comparatively small actual gain 
in heat consumption and conse- 
quently the tendency is toward 
high pressures and temperatures. 
For example, a number of recent 
installations call for a working 
pressure of 290 lb. per sq. in. abs. with superheat of 275 deg. fahr. cor- 
responding to a temperature of 590 deg. fahr. The new turbine for the 
Joliet station of the Pubhc Service Company of Northern Illinois is 
designed for 365 lb. absolute pressure with superheat of 225 deg. fahr. 
Designs are now being perfected for pressures as high as 500 pounds 
and even higher pressures have been considered. 

323. Influence of High Vacua. — The possible economy of the recipro- 
cating engine is greatly restricted by its limited range of expansion. 
Cylinders cannot be profitably designed to accommodate the rapid 
increase in the volume of steam when expanded to very low pressures. 
For example, the specific volume of 1 pound of steam under a vacuum of 



330 "S iO 



+3 
310 m 20 















/ 


\ 












/ 


\ 










si 
^1 






Va 








~/ 






^ 






I 


f/ 








V 




7 


1 


/ 








\ 








\ 


\. 










/ 










^5 


^^ 


^ 






^ 


>-' 


T^ 


"B.t.u 


Gaihgd per 


/ 


/i 


K^^i^ 




10 De 


g.Fah 


'^. 










\ 



21 



29 



25 26 27 28 
A'^acuum Barometer 30" 
80 90 100 110 120 130 140 
Temperature of Vacuum, Deg. Fahr. 



Fig. 287. Distribution of Heat during 
Adiabatic Expansion; Initial Pressure 
15 lb. Absolute; Saturated. 



STEAM TURBINES 



497 



C.10 



29 inches (referred to a 30-inch barometer) is about 667 cubic feet or 
nearly double its volume under a vacuum of 28 inches. Usually the 
exhaust is opened at a pressure of 6 or 8 pounds absolute and con- 
sequently a large proportion of the available energy is lost. The lower 
vacuum in the exhaust pipe, therefore, serves only to diminish the 
back pressure and does not affect the completeness of expansion. 
Even if it were practical to expand to 1 pound absolute, the increased 
condensation in the reciprocating engine would probably offset any 
gain due to expansion unless the steam were highly superheated. A 
study of a number of tests 
of reciprocating engines 
shows but a slight im- 
provement in overall plant 
economy due to increas- 
ing the vacuum beyond 26 
inches. Tests of steam tur- 
bines show a decrease in 
steam consumption of 
about 5 per cent for each 
inch of vacuum between 
25- and 27-inch vacuum, 
6 per cent between 27- and 
28-inch and 8 to 12 per 
cent between 28- and 29- 
inch. These values are ap-. 
proximate only since the 
influence of vacuum on the 
steam consumption varies 
greatly with the type and 
size of turbine. 

Since the volume of the steam increases very rapidly with the de- 
crease in back pressure the corresponding capacity and power required 
by the air and circulating pumps becomes proportionately larger. 
There is consequently a point where the improvement in steam economy 
fails to exceed the increased power demanded by the auxiliaries. This 
is illustrated graphically in Fig. 288. The values in Fig. 288 refer to 
a specific case only but the general principle is the same for all conditions. 
In the older types of condensing equipment the cost of maintaining the 
vacuum above 27 inches, referred to a 30-inch barometer, increased 
very rapidly with the increase in vacuum. In the modern plant vacua 
amounting to 97 per cent of the theoretical maxinmm (as determined In^ 
the temperature of the cooling water) are readily maintained with ex- 



A 
B 


Actual reduction at turbine 

Net reduction in fuel consumption 






/ 


C Equated cost of maintaining the higher vacuum 
D Final plant improvement 


/ 


















V 


/ 


















/ 


















/ 
/ 


/ 


















/ 


















/^^ 


















/ 


y 
















/ 


V 














y. 


k' 








^ 




.^^ 






<^ 






































2 


x> 


2 


6 


2 


7 


2 


8 


2d 



Fig. 288. 



Vacuum at turbine exhaust 
30 inch barometer 

Influence of Vacuum on Cost 
of Power. 



498 



STEAM POWER PLANT ENGINEERING 



cessive cost. This influence of vacuum on the economy of a 30,000-kw. 
turbo-alternator is shown in Fig. 285. 

234. Tesla Bladeless Turbine. — Fig. 289 shows a section through a 
200-horsepower experimental turbine designed by Nikola Tesla. It 
consists of a rotor composed of 25 steel disks (each ^V ii^ch thick and 
arranged on the shaft so that the length of the shaft covered by the 
disks is approximately 3.5 inches) revolving in a plain cylindrical casing. 
There are no guide plates or vanes and the viscosity and adhesion of the 
steam is depended upon for driving the rotor instead of impulse and 
reaction as in the standard type of turbine. Steam flows from the 
circumference to the center, and, when the rotor is at rest, flows by a 
short curved path, as indicated by the line in the end view, across the 




Fig. 289. Tesla Bladeless Turbine. 

face of the disk. When the rotor is up to speed the steam passes to 
the exhaust in a spiral path from 12 to 16 feet in length. Since the 
direction of rotation is determined solely by the direction of the entering 
jet it is only necessary to change the direction of the latter to effect 
complete reversal of the rotor. Mr. Tesla states that a 200-horsepower 
turbine of this type has attained a performance of 38 pounds per horse- 
power hour, initial pressure 125 pounds gauge, atmospheric exhaust, 
9000 r.p.m. (Prac. Engineer, U. S., Dec, 1911, p. 852.) The space 
occupied by this unit is only 2 feet by 3 feet and 2 feet high and the 
weight of the engine alone is 2 pounds per horsepower developed. 
This device has never been commerciahzed. 

235. "Spiro" Turbine. — Fig. 290 gives the general details of a 
high-speed rotary steam engine which has been erroneously classified 
by its builders as a turbine. It consists essentially of a pair of her- 
ringbone gears revolving in a twin cylindrical casing. Steam enters 
space a, Fig. 295, through ports pp and presses upon the gear teeth. 



STEAM TURBINES 



499 



driving them forward. The volume is increased from that indicated 
at a to that shown at h, c, d, e, and / and the energy produced is the 
product of the pressure and volume. Exhaust occurs when the ends 
of the grooves in which the action lies pass the line of contact so that 




Spiro " Turbine. 



they are no longer closed by the teeth of the opposite gear. The load 
may be varied by throttling or by cutting off the steam supply. The 
^' Spiro" is built in various sizes ranging from 1 to 200 horsepower and 
operates at 2000 to 3000 r.p.m. The following tests give an idea of the 
economy effected by this type of motor. (Power, Feb. 6, 1912, p. 188.) 



Boiler pressure, pounds gauge 

Inlet pressure, pounds gauge 

Back pressure, pounds gauge 

Horse power developed 

R.p.m 

Steam, pounds per horse-power hour 



Test 1. 



120 


130 


101.5 


115 


Atmos. 


Atmoa. 


25.3 


151 


2450 


2710 


53.2 


31.8 



Test 2. 



PROBLEMS. 

1. Steam expands adiabatically in a frictionless nozzle from an initial pressure 
of 200 lb. per sq. in. absolute, superheat 200 deg. fahr., to a back pressure of 1 in. 
absolute, weight discharged 7200 lb. per hr.; required: 

a. Velocity of the jet at the throat. 

h. Maximum spouting velocity. 

c. Diameter of the throat. 

d. Diameter of the mouth. 

e. Quality of the steam at the mouth. 

2. If the jet in Problem 1 impinges tangentially against a set of moving vanes 
and leaves them with residual velocity of 500 ft. per second, required: 

a. Velocity of the vanes, neglecting all friction and leakage losses. 

b. Horsepower imparted to the rotor. 



500 STEAM POWER PLANT ENGINEERING 

c. Pressure exerted against the vanes. 

d. Impulse efficiency of the jet. 

e. Water rate, lb. per hp-hr. 

3. Same conditions and requirements as in Problems 1 and 2 except that the 
energy-efficiency is 94 per cent and the loss of energy between inlet and exit of the 
vanes is 15 per cent. 

4. If the nozzle in Problem 1 is to be used in connection with a multi-pressure 
steam turbine, required the theoretical number of stages necessary for a peripheral 
velocity of 500 ft. per sec. Jet impinges tangentially against the rotor and all of 
the available energy is absorbed in driving the rotor. 

5. A single-stage impulse turbine (De Laval type) develops 200 hp. under the fol- 
lowing conditions: Initial pressure 153 lb. abs., back pressure 4 in. abs., superheat 
50 deg. fahr., water rate 14.4 lb. per hp-hr., nozzle angle 20 deg., peripheral velocity 
of the rotor 1200 ft. per sec. Required: 

a. Thermal efficiency. 

6. Rankine cycle ratio. 

c. B.t.u. per hp. per minute. 

6. Construct the theoretical velocity diagram for the conditions in Problem 5 
and sketch in the blade outlines. 

7. Construct the theoretical velocity diagram for a 750-hp., 2-stage Curtis turbine 
operating under the following conditions: Initial pressure 175 lb. abs., superheat 
150 deg. fahr., back pressure 2 in. abs., Rankine cycle efficiency 65 per cent, nozzle 
angle 20 degs., peripheral velocity 500 ft. per sec. Each stage consists of two ro- 
tating elements and one stationary element. 

8. Construct the velocity diagram and calculate the work done per stage in a 
frictionless reaction turbine for the following conditions: Heat drop per stage, 
16 B.t.u. per lb. of steam, peripheral velocity to be the maximum theoretically pos- 
sible for the given conditions, exit angle 30 degs., entrance angle 0. 

9. Determine the weight of water to be stored in a regenerator to operate a 1000- 
hp. exhaust steam turbine for 6 minutes if the steam supply is entirely cut off; 
pressure drop 15 to 12 lb. abs., turbine water rate 28 lb. per hp-hr. 



CHAPTER XI 

CONDENSERS 

226. General. — The primary object of condensing is the reduction 
of back pressure although the recovery of the condensate may be of 
equal importance. If a given volume of saturated steam be confined 
in a closed vessel abstraction of heat will result in condensation of part 
of the vapor with a corresponding drop in temperature and pressure. 
The greater the amount of heat abstracted the greater will be the 
amount condensed and the lower will be the temperature and pressure. 
All of the vapor can never be condensed in practice since this would 
necessitate a lowering of the temperature to absolute zero or 492 de- 
grees below the fahrenheit freezing point; consequently, the pressure 
can never be reduced to zero. With water as the cooling medium 
the minimum temperature to which the vapor can be reduced is 32 
deg. fahr. corresponding to a pressure of 0.0886 lb. per sq. in. or 0.1804 
in. of mercur}'. This represents, therefore, the lowest condenser pres- 
sure possible in practice. Condensing results in reduction of pressure 
only when the vapor is contained in a closed vessel. Thus if the vessel 
is open to the atmosphere heat abstraction will result in condensation 
but the pressure will not fall below that of the atmosphere. 

The standard atmospheric pressure at sea level and at latitude 45 
degrees is 14.6963 lb. per sq. in., corresponding to a mercury column 
29.921 inches in height, temperature of the mercur}^ 32 deg. fahr. 
For any other temperature there will be a corresponding height of 
column because of the expansion or contraction of the mercury. Steam 
tables are based on a standard pressure of 29.921 inches of mercury 
at 32 deg. fahr. and for this reason it is convenient to transfer the 
observed barometer and mercurial vacuum gauge readings to the 
32-degree standard. 

The mercury column correction for any change in temperature 
may be closely approximated by the equation 

h = h, [1 - 0.000101 ih - 0], (192) 

in which 

h = height of mercury column corrected to temperature ^, 
hi = observed height of mercury column, 
ti = observed temperature of mercury column, 
t = temperature to which column is to be referred. 

501 



502 STEAM POWER PLANT ENGINEERING 

Example 36. If the height of mercury in a vacuum gauge is 28.52 
inches, temperature 80 deg. fahr., and the barometer column is 29.85 
inches in height, temperature 62 deg. fahr., transfer the readings to 
the 32-degree standard. 

For the barometer: 

h = 29.85 [1 - 0.000101 (62 - 32)] 
= 29.77. 

For the vacuum gauge: 

h = 29.52 [1 - 0.000101 (80 - 32)] 
= 28.37. 

Absolute back pressure = 29.77 - 28.37 = 1.40. 

Vacuum referred to 32-deg. standard = 29.92 - 1.40 = 28.52. 

In condenser work it is common practice to refer the reading of the 
vacuum gauge to a 30-inch barometer, in which case it is necessary 
to increase the standard temperature of the mercury to such a figure 
as will increase the height of the barometer from 29.921 to 30 inches; 
viz., 58.15 deg. fahr. Thus, if the barometer and vacuum gauge read- 
ings are corrected to a temperature of 58.15 deg. fahr. the difference 
between the figures will give the absolute pressure in inches of mer- 
cury at 58.15 deg. fahr., and if the difference is subtracted from 30 
inches the result will give the inches of vacuum referred to a 30-inch 
barometer. According to A.S.M.E., 1915 Power Code, a 30-inch 
barometer refers in round numbers to a standard atmosphere with 
mercury at an ordinary temperature of 78 degrees. 

Example 37. Height of mercury in vacuum gauge 28.52 inches, 
temperature of mercury 80 deg. fahr., barometer 29.85 inches, tem- 
perature 42 deg. fahr.; determine the vacuum referred to a 30-inch 
barometer. 



For the vacuum gauge 



h = 28.52 [1 - 0.000101 (80 - 58.15)] 
= 28.46. 

For the barometer 

h = 29.85 [1 - 0.000101 (42 - 58.15)] 
= 29.9. 

Absolute pressure in inches of mercury at temperature 58.15 deg. 
fahr. = 29.9 - 28.46 = 1.44. 

Vacuum referred to 30-inch barometer = 30 — 1.44 = 28.56. 

According to Dal ton's Laws: (1) The mass of a given kind of vapor 
required to saturate a given space at a given temperature is the same 
whether the vapor is all by itself or associated with vaporless gases; 
(2) the maximum tension of a given kind of vapor at a given tempera- 
ture is the same whether it is all by itself or associated with vaporless 
gases; (3) in a mixture of gas and vapor the total pressure is equal to 
the sum of the partial pressures. The final pressure Pc is therefore 



CONDENSERS 503 

the combined pressure of the air Pa and that of the water vapor Py, or, 
assuming complete saturation, 

Pc = Pa + Pv. (193) 

According to the laws of Boyle and Charles the volume, pressure 
and temperature relation of an ideal gas is 

PV 

-^ = constant ( = 53.34 for dry air) (194) 

in which 

P = absolute pressure of the air, lb. per sq. ft., 
V = volume of one pound, cu. ft., 
T = absolute temperature, deg. fahr. 

Since 1 lb. per sq. ft. = 0.016 in. of mercury at 32 deg. fahr., equation 
(194) may be conveniently expressed 

^ = 0.755 (195) 

in which 

Pa = absolute pressure, in. of mercury. 

By means of equations (193) and (195) all problems involving a satu- 
rated mixture of air and water vapor may be readily solved. See 
Chapter XXV for a discussion of the properties of dry, saturated and 
partially saturated air. 

\' Example 38. If the absolute pressure in a condenser is 4 inches of 
mercury and the temperature of the air-vapor mixture is 100 deg. fahr., 
required the percentages of air by weight in the mixture. 

From steam tables the pressure of vapor corresponding to a tem- 
perature of 100 deg. fahr. is 1.93 inches of mercury. 

Hence, from equation (193), 

Pc = Pa-\- Pv, 

4 = P„ + 1.93, 
Pa = 2.07. 

Let V = volume of the condenser chamber, cubic feet. 
Then 0.00285 v = weight of vapor in the chamber (0.00285 = den- 
sity of water vapor at 100 deg. fahr.), and 

0.08635 X ^^ X ^II^^^qq V = 0.00491 v = weight of dry air in the 

chamber. (0.08635 = density of air at deg. fahr. and 29.92 inches of 
mercury pressure.) 

The total weight of the mixture is 

0.00285 V + 0.00491 v = 0.00776 v, 

and the percentage of air in the mixture is 

^•^t 7A "" =" 0.632 or 63.2 per cent. 
0.00776 V 



504 



STEAM POWER PLANT ENGINEERING 



Curve A, Fig. 291, shows the influence of vacuum on the percentage 
of air in the air- vapor mixture for a constant air pressure of 0.1 in. 

Curve B, Fig. 291, shows the difference between the temperature of 

saturated vapor correspond- 
ing to the total pressure in 
the condenser to that of the 
actual vapor for various 
vacua with constant air pres- 
sure of 0.1 in. 

For data pertaining to the 
amount of air carried into 
condensers see paragraphs 
306-9. 

Example 39. If the tem- 
perature within a condenser 
is 110 deg. fahr. and there 
is entrained with the steam 
0.2 of a pound of air per 
pound of steam, required the 



23 

22 

21 

20 

19 

18 

cl7 

OIG 

fel5 

14 

13 

12 

11 

10 

9 

8 





























( 




































/ 




















Pe 


rCen 


b> 


W 


eight 


of Air 


k 


/ 




















to 


St* 


an 


P 


us 


Ail 


M 


ixt 


ire 


/ 










( 






























( 






































/ 










1 




























/ 












/ 


























J 


/ 










/ 


/ 


























/ 


'T 


Jtn 


jer 


atu 


re 


/ 


























/ 




D 


iff 


;re 


ice 


>" 


























/ 










y 


y 
























> 


/ 








U- 


r" 


























r^. 




L 


-A 


^ 






Note:-Partial Air Pressure 
Constant =0.1 In. llercury 




— ^ 


^ 


^ 
















^ 


^ 


















1 1 ! M 1 1 1 



28 28.a 28.4 28.6 28.8 29 29.2 29.4 29.5 
" Vacuum Referred to 30 In. Barometer 



Fig. 291. Percentage of Air in Mixture and 
Difference of Temperatures Corresponding to 
Total Pressure and that Actually Existing, for maximum degree of vacuum 
Constant Air Pressure of 0.1 in. obtainable. 

One pound of saturated 
steam at a temperature of 110 deg. fahr. occupies a volume of 265.5 
cu. ft. The corresponding vapor tension is 2.589 in. of mercury. This 
must also be the volume occupied by 0.2 pound of air mixed with it 
and the temperature of the air is that of the vapor (110 deg. fahr.). 
Then from equation (195), 

p 0.755 X (110 + 46 0) „ __ . , 

Pa = 2(^5 5 — 02 ^ 0.322 m. of mercury. 

From equation (194) 

Pc = Pa + Pv 

= 0.322 + 2.589 = 2.911 in. of mercury. 

And the vacuum 

= 29.921 - 2.911 = 27.01 in. of mercury. 

If no air were present the maximum vacuum would be 

29.921 - 2.589 = 27.332 in. of mercury. 

The lower the temperature of the vapor the greater will be the in- 
fluence of the air, thus, if the temperature in the preceding problem 
were 80 deg. fahr. the pressure of the air would be 0.306 and that of 
the vapor would be 0.505. The ill effects from air entrainment at 
low vacua are apparent. 

Air in Condensers: Power, Feb. 29, 1916, p. 291; June 13, 1916, p. 834, Mar. 14, 
1916, p. 376: Elec. Wld., July 8, 1916, p. 84; Jour. A.S.M.E., Feb., 1916, p. 190. 

A condenser is a device in which the process of condensation and 
subsequent removal of the air and condensed steam is continuous, the 



CONDENSERS 505 

degree of vacuum obtained depending upon the tightness of valves and 
joints, the quantity of entrained air, and the temperature to which the 
condensed steam is reduced. 

The degree of vacuum may be expressed in different ways. (1) Ex- 
cess of the atmospheric pressure over the observed vacuum. For 
example, a 26-inch vacuum implies that the pressure of the atmosphere 
is 26 inches of mercury above the pressure in the condenser. (2) Per 
cent of vacuum, by which is meant the ratio of the observed vacuum to 
the atmospheric pressure. Thus, with the barometer standing at 30 
inches, a vacuum of 26 inches may be expressed as 100 X §f = 86.6 
per cent vacuum. This method of expression gives an idea of the 
efficiency of the condensing system. For example, the degree of 
vacuum indicated by 26 inches would be 93 per cent with a barometric 
pressure of 28 inches but only 84 per cent when the barometer reads 
31 inches. (3) Absolute pressure. Thus a 26-inch vacuum referred to 
a 30-inch barometer would be indicated as a pressure of 30 — 26 = 4 
inches absolute, or 1.99 pounds per square inch. 

The place of measurement of the vacuum should be stated since the 
lowest back pressure will be found at the air-pump suction, a higher 
pressure in the body of the condenser and the highest at the prime mover 
exhaust nozzle. 

227. Effect of Aqueous Vapor upon the Degree of Vacuum. — The 

futility of attempting to better the vacuum by exhausting the vapor is 
best illustrated by a specific example. 

Example 40. Required the volume of aqueous vapor to be with- 
drawn per hour from a condenser operating under the following condi- 
tions, in order that the vacuum may be increased one pound per square 
inch: Temperature of discharge water 125 degrees; corresponding 
vapor tension 4 inches of mercury; barometer 30 inches; relative 
vacuum 26 inches; horsepower 100; steam consumption 20 pounds per 
horsepower-hour; cooling water 25 pounds per pound of steam condensed. 

100 X 20 X 25 = 50,000 pounds of cooling water per hour. 
= 833 pounds of cooling water per minute. 

Now to increase the vacuum one pound per square inch, approxi- 
mately 2 inches of mercury, the temperature of the water must be 
lowered to 102 deg. fahr., that is, 833 (125 - 102) = 19,159 B.t.u. 

19 159 
must be abstracted from the water in one minute, or ' = 18.6 

iUoU 

pounds of water to be evaporated per minute. (1030 = average heat 
of vaporization of water under 26 to 28 inches of vacuum.) Now, one 
pound of vapor at 102 to 125 deg. fahr. has an average volume of 270 
cubic feet. 

Therefore 18.6 X 270 = 5022 cubic feet of vapor must be exhausted 
per minute to increase the vacuum from 26 to 28 inches, which while 
not impossible is manifestly impracticable for condenser practice. 

In the Westinghouse-Leblanc refrigerating system cooling is effected 
by the withdrawal of aqueous vapor by means of an air pump. 



506 



STEAM POWER PLANT ENGINEERING 



228. Gain in Power due to Condensing. — The advantages to be 
gained by decreasing back pressure may be most readily illustrated 
by the following example: 

Example 4\. A non-condensing engine taking steam at a pressure 
of 100 pounds absolute and cutting off at one-quarter stroke will have, 
theoretically, a mean effective pressure on the piston of 44.6 pounds 
per square inch, the back pressure being 14.7 pounds per square inch 
absolute. If the engine exhausts into a condenser against a 26.5-inch 
vacuum (1.7 pounds absolute) the mean effective pressure will be in- 
creased to 44.6 + (14.7 — 1.7) = 57.6 pounds per square inch, resulting 
in a gain in power which may be expressed 



Hp. = 



PrAS 

33,000 ' 



(196) 



in which 

Up. = horsepower gained, 

Pr = reduction in back pressure, pounds per square inch, 
A = area of the piston in square inches, 
S = piston speed in feet per minute. 

If P = mean effective pressure on the piston when running non-con- 
densing, the percentage of increase of power may be expressed 

Percent = 100 ^^ (197) 

In the above example the percentage of power gained would be 

100 -^ = 29.2 per cent. 
44.6 

The actual gain due to the use of the condenser would be much 
less than this, depending upon the type of engine and conditions of 
operation. 

TABLE 90. 

PRESSURE OF AQUEOUS VAPOR IN INCHES OF MERCURY FOR EACH DEGREE F. 

(Marks and Davis.) 



30° 

40° 

50° 

60° 

70° 

80° 

90° 

100° 

110° 

120° 

130° 

140° 



.248 
.362 
.522 
.739 
1.03 
1.42 
1.93 
2.60 
3.44 
4.52 
5.88 



.257 
.376 
.541 
.764 
1.06 
1.46 
1.98 
2.66 
3.53 
4.64 
6.03 



.180 
.268 
.390 
.560 
.790 
1.10 
1.51 
2.04 
2.74 
3.63 
4.76 
6.18 



3° 



.188 
.278 
.405 
.580 
.817 
1.13 
1.55 
2.11 
2.82 
3.74 
4.89 
6.34 



.195 
.289 
.420 
.601 
.845 
1.17 
1.60 
2.17 
2.90 
3.84 
5.02 
6.51 



203 

300 

436 

622 

873 

21 

65 

24 

99 



3.95 
5.16 
6.67 



.212 
.312 
.452 
.644 
.903 
1.25 
1.71 
2.30 
3.07 
4.06 
5.29 
6.84 



.220 
.324 
.468 
.667 
.964 
1.30 
1.76 
2.37 
3.16 
4.17 
5.43 
7.02 



.229 
,336 
.486 
.690 
,996 
,33 
,81 
44 



3.25 

4.28 
5.58 
7.20 



.238 
.349 
.50^ 
.714 
1.03 
1.37 
1.87 
2.51 
3.34 
4.40 
5.73 
7.38 



CONDENSERS 



507 



With steam turbines the advantage gained by reduction of back 
pressure is more marked than with the reciprocating engine, though 
theoretically the same for the same range of expansion. Initial con- 
densation, leakage past valves, and other sources of loss prevent a 
reciprocating engine from benefiting from a good vacuum to the same 
extent as a turbine. See paragraph 223. 

Referring again to the example given above, if the steam is cut off at 
about one-sixth stroke, the work done when running condensing will be 
the same as when running non-condensing and cutting off at one-quarter. 
Theoretically the steam consumption will be decreased nearly in pro- 
portion to the reduction in cut-off. Generally speaking, a condensing 
engine will require from 20 to 30 per cent less steam for a given power 
than a non-condensing engine. (See results of engine tests, paragraph 
181.) This decrease in steam consumption is only an apparent one. 
If steam is used by the auxiliaries in creating the vacuum, the amount 
must be added to that consumed by the engine, unless the steam ex- 
hausted by the former is utihzed to warm the feed water, in which case 
only the difference between the heat entering the auxiliaries and that 
returned to the heater should be charged against the engine. The power 
necessary to operate the condenser auxiliaries varies from one to six 
per cent of the main engine power, depending upon the type and con- 
ditions of operation. 

In power plants where the exhaust steam is not used for heating or 
manufacturing purposes, the engines are almost invariably operated 
condensing, provided there is an abundant supply of cooling water. 
Even if the water supply is limited, it is often found to be economical 
to use some artificial cooHng device, notwithstanding the high first cost 
and cost of operation of the latter. 

Some of the considerations affecting the propriety of running condens- 
ing and the choice of condensing systems are taken up in paragraph 249. 

229. Classiflcatlon of Condensers. — The following is a classification 
of a few well-known condensers: 

Standard low ^ Sl^^'"'^*''''' 



1. Jet condensers. , 



Parallel current (a). 



Siphon !EST"*- 



Ejector. 



Schutte. 
Korting. 



Counter current (6) 



(Barometric | ^g^g,,. 

■ j ( LeBlanc. 

( High vacuum • • ] Wheeler. 

( Worthington. 

C Single-flow Barag\vanath. 

Water cooled (a) < Double-flow Wheeler. 

( Multi-flow Wainwright. 

Air cooled (h) ( Forced draft Fouche. 

I Natural draft Fennel 1. 

Evaporative (c) Ledward. 

Condensers may be divided into two general groups: 

1. Jet condensers, in which the steam and cooling water mingle and 

the steam is condensed by direct contact, Figs. 292 to 300. 



2. Surface condensers 



508 STEAM POWER PLANT ENGINEERING 

2. Surface, condenser Sj in which the steam and coohng medium are 
in separate chambers and the heat is abstracted from the steam by con- 
duction, Figs. 305 to 309. 

Jet condensers may be further grouped into two classes, according to 
the direction of flow of the air and cooling water: 

(a) Parallel-current condensers^ in which the condensed steam, cool- 
ing water, and air flow in the same direction, collect at the bottom of 
the condenser chamber, and are exhausted by a suitable pump, Fig. 292. 

(6) Counter -current condensers, in which the cooling water and con- 
densed steam flow from the bottom of the chamber, while the air is 
drawn off at the top. Fig. 301. 

Parallel-current condensers may be subdivided into three classes: 

(1) Standard condensers, in which the cooling water, condensed steam, 
and air are exhausted by a vacuum pump. Fig. 292. 

(2) Siphon condensers, in which the coohng water, condensed steam, 
and air are exhausted by a barometric column. Fig. 297. 

(3) Ejector condensers, in which the condensed steam and air are 
exhausted by the coohng water on the ejector principle, i^ig. 298. 

Surface condensers may be classified according to the nature of the 
cooling medium as 

(a) Water-cooled condensers. 

(6) Air-cooled-condensers. 

(c) Evaporative condensers, in which the condensation of the steam is 
brought about by the evaporation of a fine stream of water trickhng on 
the surface of the tubes. 

330. Standard Low-level Jet Condensers. — Fig. 292 shows a sec- 
tion through a Worthington jet condenser, illustrating the low-level 
type in which the condensing water is drawn into the apparatus by the 
vacuum. When the pump is started a partial vacuum is created in 
the suction chamber above the valves H, H in the cone F. As soon as 
sufficient air has been exhausted, cooling water enters at B with a 
velocity depending upon the degree of vacuum in chamber F and the 
suction head, and is divided into a fine spray by the adjustable serrated 
cone D. The spray mingles with the exhaust steam entering at A 
and both move downwards with diverse velocities. The steam gives 
up its heat to the water and condenses. The velocity of the steam 
diminishes in its downward path to zero, while the velocity of the water 
increases according to the laws of falhng bodies. The condensed steam, 
cooling water, and air collect at the lower part of the condenser and are 
exhausted by the wet air pump G, from which they are forced through 
opening J to the hot well. The vacuum in chamber F will depend upon 
the vapor tension of the warm water in the bottom of the well, the 



CONDENSERS 



509 



amount of air carried along by the cooling water and steam, and the 
tightness of valves and joints. In case the water accumulates in 
the condenser cone F, either by reason of an increased supply or by a 
sluggishness or even stoppage of the pump, the condensing surface is 




Fig. 292. Worthington Independent Jet Condenser. 

reduced to a minimum, as soon as the level of the water reaches the 
spray pipe and the spray becomes submerged, and only a small annular 
surface of water is exposed to the exhaust steam. The vacuum is 
immediately broken, and the exhaust steam escapes by blowing through 
the injection pipe and through the valves of the pump and out the dis- 
charge pipe at J, forcing the water ahead of it; consequently flooding of 
the steam cylinder cannot occur. In starting up the condenser a partial 
vacuum for inducing a flow of injection water into the condenser cham- 



510 



STEAM POWER PLANT ENGINEERING 



ber may be created by the pump if the suction lift is not too great. 
Many engineers, however, prefer to install a small forced injection or 
priming pipe the function of which is to condense sufficient steam to 
produce the necessary partial vacuum. 

Fig. 293 shows a section through the condensing chamber and air 
pump of a Blake vertical jet condenser with an automatic vacuum- 
breaking device. The injection water enters at opening marked '^injec- 
tion" and flows through the adjustable ''spray" nozzle in a fine spray, 



steam 
Cylinder 




Air Pump 

Fig. 293. Section through a Blake Jet Condenser. 

at an angle of about 45 degrees, and impinges on the conical sides of the 
upper condenser chamber. The spray fafis from the sides to the pro- 
jecting ledges shown in the illustration. The ledges prevent the spray 
from falhng directly to the bottom of the chamber and insure an efficient 
mingUng of steam and cooling water. A perforated copper plate is 
substituted for the shelves when the force of the injection water is not 
sufficient to produce spray. The circulating water and condensed 
steam together with the non-condensable gases are drawn off at the 



CONDENSERS 511 

bottom of the chamber. The vacuum-breaking device is shown at the 
right of the figure. When the rising water reaches the level of the float 
chamber, as in the case of an accidental stoppage of the air pumps, the 
float is raised and forces a check valve from its seat and allows an inrush 
of air to break the vacuum, thus preventing further suction of water 
into the condenser and consequent flooding of the engine. A is the 
forced injection or ''priming" inlet used in starting up when the suction 
lift is considerable. 

231. Injection Orifice. — The velocity of water entering a jet con- 
denser, neglecting friction, may be determined from the equation 

V = V2gh, (198) 

where 

V = velocity of the water in feet per second, 
g = acceleration of gravity = 32.2, 
h = total head in feet. 

If p = pressure below the atmosphere in pounds per square inch, 
hi = distance in feet between the source of supply and the injection 
orifice, 
then h = 2.Sp zt hi, (199) 

and equation (198) may be written 

V = 8.025 V2.3 p db hi. (200) 

If the supply is under pressure, h is positive; if under suction, it is 
negative. 

Example 42. What is the theoretical velocity of water entering a con- 
denser with 26-inch vacuum (referred to 30-inch barometer); suction 
head 8 feet? 

Here p = pressure in pounds per square inch, corresponding to 26 
inches of mercury = 12.8 pounds per square inch. 

hi = 8. 

V = 8.025 V2.3 X 12.8 - 8 
= 37.1 feet per second 
= 2226 feet per minute. 

In proportioning the injection orifice in practice the maximum 
velocity of flow is assumed to be between 1500 and 1800 feet per minute, 
or, approximately, area of injection orifice in square inches = weight of 
injection water in pounds -^ 650 to 780. (''Manual of Marine Engineer- 
ing, " Seaton, p. 204.) A rough rule gives area of orifice = area of low- 
pressure piston in square inches -^ 250. (Seaton, p. 204.) 

332. Volume of the Condenser Cliamber. — According to Thurston 
the volume of a jet' condenser should be from' one fourth to one half 
that of the low-pressure engine cylinder. ("Steam Engine Manual," 
Thurston, II, 127.) 



512 



STEAM POWER PLANT ENGINEERING 



According to Hutton the volume should not be less than that of the 
air pump and should approximate three fourths that of the engine 
cylinder in communication with it. 

233. Injection and Discharge Pipes. — In practice the diameter of 
the injection pipe is based on a velocity of 400 to 600 feet per minute 
and that of the discharge pipe of 200 to 400 feet per minute; the lower 
figures for pipes under 8 inches in diameter, the upper range for larger 
diameters. 

(Atmospheric relief valves. — See paragraph 363.) 

334. High-vacuum Jet Condensers. — The standard low-level jet con- 
denser is not suitable for high vacua because of the hmited air capacity 



-Watei;, 




SECTION M.M 
THROUGH AIR PUMP 



Fig. 294. Westinghouse-Leblanc Multi-jet High- vacuum Condenser System. 



of the combined air and circulating-water pump. Even with a tight 
system considerable air is carried into the condenser with the circu- 
lating water and efficient removal of the air necessitates a larger pump 
capacity than is usually furnished with this type of condenser. Low- 
level jet condensers may be operated with a high degree of vacuum by 



CONDENSERS 



513 



equipping them with independent air and circulating pumps. Ex- 
amples of this type of jet condenser are illustrated in Figs. 294 to 296. 
Referring to Fig. 294 which gives several views of the Leblanc type of 
condenser, steam enters the condensing chamber as indicated and meets 
the coohng water injected through spray nozzle C. The condensed 
steam and injection water fall to the bottom of the vessel and are re- 
moved by centrifugal pump M. Air saturated with water vapor is with- 
drawn by centrifugal air pump P through suction opening 0. Re- 
ferring to section NN through the air pump it will be seen that this 
device consists primarily of a reverse Pelton wheel in conjunction 



-■:i3i::=rj 




Fig. 295. C. H. Wheeler Low-level, High-vacuum Jet Condenser. 

with an ejector. Sealing water is introduced through the branch 
indicated by dotted outline into the central chamber G from which 
it passes through port H. It is then caught up by the blades P of the 
Pelton wheel, which is rotated at a suitable speed, and ejected into the 
discharge cone in the form of thin sheets having a high velocity. These 
sheets of water meet the sides of the discharge cone and thus form a 
series of water pistons, each of which entraps a small pocket of air and 
forces it out against the atmospheric pressure. In passing through the 
air pump the sealing water receives practically no increase in temperature, 
hence the same water may be used over and over again. The air pump 
rotor and main pump runner are enclosed in a common casing mounted 
on the same shaft. There is a clear passage through the condenser 
and pump, so that, should the pump stop for any reason, air rushes into 



514 



STEAM POWER PLANT ENGINEERING 



the condenser through the air pump and immediately breaks the vacuum. 
In starting up the condenser steam is turned into auxiUary nozzle L, 
section NN, for a few moments, thus creating sufficient vacuum to 
start the regular flow of water through the air pump. 

Any type of air pump may be used in connection with a suitable 
circulating pump but the majority of low-head, high-vacuum jet con- 
densers are equipped with the hydraulic type. Recent experiments 



Exhaust Steam 




Fig. 296. Wheeler Low-level Centrifugal Jet Condenser. 



indicate that the steam jet type of air ejector may supersede the 
mechanically-operated air pump within a very short time. (See para- 
graph 309.) 

235. Siphon Condensers. — Fig. 297 shows a section through a 
Baragwanath siphon condenser, illustrating the principles of a parallel- 
current barometric condenser. The cooling water enters the side of 
the condenser chamber at A and passes downward in a thin annular 
sheet around the hollow cone D. The exhaust steam enters at B and 
is given a downward direction by the goose neck C. It flows through 
the nozzle Z) and is condensed within the hollow cone of moving water, 
the combined mass including the entrained air discharging through the 
contracted throat E at high velocity into the tail pipe 0. The water 
column in the tail pipe must be enough to overcome the pressure of the 



CONDENSERS 



515 



atmosphere; i.e., it should be 34 feet or more above the surface of the 
hot well, otherwise water would rise within this pipe to a height cor- 
responding to that of the barometer, which is approximately 34 feet for 
a barometric pressure of 30 inches of mercury. This is not strictly 
true when the condenser is in full opera- 
tion, as the injector effect of the moving 
mass is sufficient to overcome several 
pounds pressure, and the tail pipe may 
be less than 34 feet, but to provide against 
any possibiUty of the water being drawn 
into the cylinder of the engine the length 
is made greater than 34 feet. The spray 
cone D is adjustable and admits of close 
regulation of the water supply without 
changing the annular form of the stream. 
The condensing water may be supphed 
under pressure or under suction. For 
lifts not greater than 15 feet no supply 
pump is necessary, the water being raised 
by the siphon action of the condenser. 
This condenser requires the same amount 
of cooling water per pound of steam as 
the standard jet condenser, and is capa- 
ble of maintaining a vacuum of from 24 
to 27 inches. A vacuum of 28J inches 
has been recorded for a condenser of this 
general type. (Trans. A.S. M.E., 26-388.) 
An atmospheric reUef valve G is provided in case the vacuum fails from 
any cause, which will permit the steam to escape to the atmosphere. 

The above type of condenser is adapted to very muddy cooling 
water, since no filtration is necessary beyond the removal of such solid 
matter as may clog up the annular space H. 




Fig. 



297. Baragwanath Siphon 
Condenser. 



Siphon Condensers, Discussion: Trans. A.S.M.E., Vol. 26, p. 388. Siphon Con- 
densers: Electrical World, June, 1897, p. 818; Engr. U. S., Jan., 1906. 



236. Size of Siphon Condensers. — The size of siphon is indicated by 
the diameter of the engine exhaust pipe. 

Table 91 gives the sizes of barometric condensers as manufactured 
by prominent makers. 

The diameter of the throat may be closely approximated by the 
empirical formula 

Diam. in inches = 0.0077 VWw, (201) 



516 



STEAM POWER PLANT ENOINEERING 



in which 

W = weight of steam to be condensed per hour, 
w = weight of water required to condense one pound of steam. 

The maximum width of the annular opening for the admission of 
water may be obtained from the empirical formula 



Width in inches = 



Ww 



39,550 6^' 
in which 

d = diameter of the nozzle or bottom of the cone in inches. 
W and w as in equation (201). 

TABLE 91. 
SIZE OF SIPHON CONDENSERS. 



(202) 



steam to be Condensed, 


Size Usually 

Furnished, 

Inclies. 


Steam to be Condensed. 


Size Usually- 


Pounds per 
Hour. 


Pounds per 
Minute. 


Pounds per Hour. 


Pounds per 
Minute. 


Furnished, 
Inches. 


2,000 
3,000 
4,000 
5,000 
6,000 


33 
50 
60 
83 
100 


5 

7 
8 
9 
9 


8,000 
10,000 
15,000 
20,000 


133 
166 
250 
333 


10 
12 
14 
14 



Vacuum 26 inches; barometer 30 inches. 

237. Ejector Condenser. — Fig. 298 shows a section through a Schutte 
exhaust steam ''induction" condenser, illustrating the principles of the 
ejector condenser in which the momentum of flowing water ejects the 
discharge without the aid of the circulating pump. Exhaust steam 
enters the ejector through the opening marked ''exhaust," passes 
through a series of inclined orifices and nozzles at considerable velocity, 
and, meeting the cooling water in the inner annular chamber, is con- 
densed. The cooling water is drawn in continuously through the opening 
marked "water, " by virtue of the vacuum formed, and sufficient velocity 
is imparted to the jet to discharge the combined mass, of condensed 
steam, cooling water, and air against the pressure of the atmosphere. 

The condenser should be installed vertically with three feet of pipe 
between the strainer and the head of the condenser and should be ar- 
ranged as shown in Fig. 299. There should be a clear discharge of not 
less than two feet below the bottom flange of the apparatus to the level 
of Ihe water in the discharge sump, or hot well. It is advisable that 
the end of the discharge pipe be sealed under water, unless there is a 



CONDENSERS 



517 



horizontal discharge main, and trap to water seal at the bend immedi- 
ately under the condenser. Except with condenser of very large size 
a difference of level between supply and discharge of 30 feet will usually 
give the necessary pressure of water at the condenser with full allow- 




Exhaust 



Strainer 



/?B^ 




Discharge 
Fig. 299. Piping for Schutte Ejector Condonser. 



lEIxhaust to Condenser 



Air Pipe 



-Vacuum Breaker 




►Discharge 



Discharge 

Fig. 298. Schutte 
Ejector Condenser. 



Thermometer 
Connection 



Circulating Water 

Fig. 300. Section through Condensing Chambers of Kort- 
ing Multi-jet Condenser. Chamber Capable of Main- 
taining a Vacuum of 95 Per Cent of the Ideal without 
the Use of Air Pumps. 



ance for friction losses. These condensers are made in all sizes conform- 
ing with exhaust pipe diameters of li to 24 inches. The same amount 
of cooling water is required as for jet condensing and vacua of 20 to 
25 inches are readily obtained. 

238. Barometric Condensers.* — Fig. 301 shows a section through 
a Weiss counter-current condenser, illustrating the principles of a 
barometric jet condenser. The cooling water enters the upper part 

* The author has been informed that the word "Barometric" in connection with 
jet condensers is the registered trade mark of the Albergcr Condenser Company. 



518 



STEAM POWER PLANT ENGINEERING 



of the condensing chamber A through pipe A^ and falls in cascades, as 
shown in the figure, to tail pipe B, from which it flows by gravity to 
the hot well. The exhaust steam enters chamber A through pipe D, 
and, coming in contact with the cold-water spray, is condensed. The 
air is exhausted from the top of the condenser by a dry vacuum pump 
through pipe F. In flowing to the pump the air passes upwards through 

the water spray and its temperature is 
lowered to that of the injection water, 
thereby reducing the volume to be ex- 
hausted. Any moisture passing over 
with the air is separated at G before 
reaching the air pump, and flows out 
through the small barometric tube H. 
The cooling water is forced to the 
condenser chamber through pipe N by 
any positive displacement pump, the 
actual head pumped against being the 
difference between the total height and 
that of a column of water correspond- 
ing to the degree of vacuum in the 
l_y 1^^^ ~-^^ i r condenser. The main barometric tube 

T^ -% ;~^..i J^ ''^f or tail pipe B through which the water 

is discharged is 34 feet or more in 
length and is provided with a foot 
valve C. The counter-current prin- 
ciple permits a much higher tempera- 
ture of hot well for the same degree 
of vacuum than does the parallel cur- 
rent, a hot- well temperature of 120 de- 
grees and a vacuum of 27 inches being 
readily maintained. A small pipe K 
connecting the main condenser with 
the small barometric tube H insures at 
all times a sufficient quantity of water in the small auxiliary hot well 
to seal the tube. The water from this auxiUary hot well flows over a 
weir, as indicated, into a counterweighted bucket M, the latter having 
a hole in the bottom which allows the normal flow to escape. But in 
case a sudden heavy overload is thrown on the engines, and the ad- 
justment is for a fight load, the temperature of the discharge will reach 
the boifing point and an abnormal quantity of water will flow down 
the small barometric tube. This wiU cause the water to flow into the 
bucket much faster than the opening in the bottom can dispose of it; 




Fig. 301. 



Weiss Counter-current 
Condenser. 



CONDENSERS 



519 



as a result the bucket will increase in weight and will open up a free- 
air valve L which reduces the vacuum two or three inches and raises 
the boiling point without '^dropping" the vacuum entirely. E is the 
atmospheric relief valve. 

Fig. 302 shows a section through the condensing chamber of an 
Alberger barometric condenser. In principles of operation the con- 
denser is similar to the Weiss, but differs considerably in details. Ex- 
haust steam enters at A and divides into two streams, one flowing 
directly to the inner chamber D, the other through the annular space E. 
Cooling water enters through 
B and is broken up into a fine 
spray by the serrated cone F, 
which is hung upon a long 
spring, thus automatically ad- 
justing itself to the quantity of 
water entering the condenser. 
After condensing the exhaust 
steam in the inner cylinder the 
partly heated spray of cooling 
water in falling is brought in 
contact with the exhaust steam 
which enters through the an- 
nular space. This process per- 
mits of a high hot-well tem- 
perature without affecting the 
degree of vacuum. The air 
which is not entrained by the 
cooling water and carried down 
the tail pipe collects under 
the spray cone F and ascends 




Fig. 302. Section through Condensing Cham- 
ber, Alberger Barometric Condenser. 



through the tubular support of the cone into the air cooler. This 
air cooler is simply a small chamber in which the non-condensable 
gases are cooled by a small portion of the circulating water before 
they are withdrawn by the air pump. The circulating water used for 
the purpose is forced into the cooling chamber through pipe K and falls 
through serrated openings in the bottom to the condenser proper. 
The air enters the chamber through these same openings, and is with- 
drawn by the air pump. Surrounding the cooler is a separating space 
of large capacity to allow the subsidence of any entrained moisture 
before the air reaches the vacuum pump. 

Fig. 303 shows a section through a Tomlinson type B barometric 
condenser which differs from the conventional type in the addition of an 



520 



STEAM POWER PLANT ENGINEERING 



overflow or auxiliary tail pipe. The main tail pipe takes care of the 
light loads and the overflow comes into service only on full loads and 
overloads. This arrangement reduces the quantity of circulating 
water required at light loads since it is not necessary to keep a large 

tail pipe filled with water as is the 
case with the single pipe design. 

;339. Condensing Water: Jet Con- 
densers. — In a jet condenser the 
coohng water and exhaust steam 
mingle, and the degree of vacuum is 
a function of the final or discharge 
temperature; thus the quantity of 
cooling water required depends upon 
its initial temperature, the tempera- 
ture of the discharge water, and the 
total heat of the steam entering the 
condenser. If the steam in the low- 
pressure cylinder at exhaust is dry 
and saturated, and there is no air 
entrainment the heat entering the 
condenser will correspond to the 
total heat of saturated steam at con- 
denser pressure. This condition is 
not likely to occur in practice since 
exhaust steam usually carries con- 
siderable moisture and there will be 
more or less air entrained with it. 
Furthermore, the cooling water con- 
tains air in varying amounts so that 
the total amount of air entering 
the condenser may be considerable. 
Neglecting radiation and leakage the heat absorbed by the cooling 
medium must be equal to that given up by the steam and its air en- 
trainment. The heat exchange may be expressed 




Fig. 



303. Tomlinson Type B Baro- 
metric Condenser. 



R = 



Hm — 



(1-2 



(203) 



in which ^2 — Qo 

R = weight of injection water necessary to condense and cool one 
pound of air-vapor mixture, 
Hm = heat content of the air-vapor mixture at condenser pressure, 
B.t.u. per lb. above 32 deg. fahr., 
^2 = heat of liquid of the discharge water, B.t.u. per lb., 
Qo = heat of liquid of the injection water, B.t.u. per lb. 



CONDENSERS 521 

In practice it is sufficiently accurate to neglect the influence of the 
air on the heat content of the exhaust steam and circulating water, 
and the mean specific heat of water under condenser conditions may be 
taken as unity so that equation (203) may be written, 

B = H-h + ^2 ^ ^204) 

in which 

H = heat content of the exhaust, B.t.u. per lb. above 32 deg. fahr., 
ti = temperature of the discharge water, deg. fahr., 
ti = temperature of the injection water, deg. fahr. 
It has been shown (paragraph 177) that 

w 

Wi 

in which 

Hi = initial heat content of the steam entering the prime mover, 

B.t.u. per lb. above 32 deg. fahr., 
Hr = heat lost by radiation from the prime mover and exhaust piping, 

B.t.u. per lb. of steam admitted, 
w = water rate, lb. per brake hp-hr., 
Wi = water rate, lb. per kw-hr. 

In a well-lagged piston engine with short connection to the con- 
denser the loss by radiation varies from 0.3 to 2.0 per cent, but seldom 
exceeds 1 per cent of the total heat admitted, and in a turbine this 
loss is even less, and 0.5 per cent is a very liberal allowance. The 
temperature of the discharge water will approximate that of the vapor 
at its partial pressure. For air-free steam this will correspond to that 
of vapor at total condenser pressure. In high-vacuum jet condensers 
in which the air' pressure is kept very low this depression of the hot- 
well temperature will range from to 5 degrees below that of vapor 
at total condenser pressure, and in the ordinary low-vacuum condenser 
it may range from 15 to 25 degrees below. The influence of air en- 
trainment for a specific case is illustrated in Fig. 291. The minimum 
weight of cooling water for air-free steam at various vacua is shown 
graphically in Fig. 304. 

Example 43: Determine the amount of cooling water necessary per 
pound of steam for a standard low-vacuum jet condenser operating 
under the following conditions: Engine uses 16 lb. steam per brake 
hp-hr., initial pressure 140 lb. per sq. in. absolute, superheat 50 deg. 
fahr., vacuum 26 inches referred to a. 30-inch barometer, temperature 
of injection water 70 deg. fahr. 



522 



STEAM POWER PLANT ENGINEERING 



Fig. 304. 



40 



50 



60 70 80 

Temperature Degrees Fahrenheit 



90 



mo 



TOIV 








































































































rJU 




















































180 
































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f, 














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^^ 






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^^ 


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^ 












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=s 


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110 



Curves showing Minimum Ratio of Circulating Water to Steam 
Condensed for Various Initial Temperatures. 



From steam tables Hi = 1221; assume Hr = 1 per cent of Hi, then 

2546 



H = 1221 - 0.01 (1221) 



16 



= 1050. 



The temperature ts of vapor corresponding to an absolute pressure 
of 4 inches = 126 deg. fahr.* Assume ^2 = ^s — 15 = 111. 

* This is not the actual temperature in the condenser. The actual^ temperature 
will be that corresponding to the partial pressure of the vapor. For convenience in 
calculation the temperature in the condenser is assumed to correspond to that of the 
total pressure and the temperature depression of the hot well is then based on this 
hypothetical temperature. When the extent of air leakage and entrainment is 
known the actual temperature in the condenser may be readily calculated. 



CONDENSERS 



523 



Substituting these values in equation (204), 
1050- 111 +32 



R 



111 - 70 



= 24.3 lb. 



Example 44. Determine the amount of cooling water necessary per 
pound of steam for a high- vacuum jet condenser operating under the 
following conditions. Turbine uses 14 lb. steam per kw-hr., initial 
pressure 165 lb. per sq. in. absolute, superheat 120 deg. fahr., vacuum 
29 inches referred to a 30-inch barometer, temperature of injection 
water 65 deg. fahr. 

From steam tables, Hi = 1262; ts = 79; assume Hr = 0.005 H,-. 
(This is so small that it may be omitted, particularly in view of other 
assumptions which may be made.) 



Then, 



Assume 



H = 1262 



0.005 (1262) - ^^ = 1012. 



R 



t2-ts-4: 

1012 -75+32 



14 
75 



= 96.9 lb. 



340. Water-cooled Sur- 
face Condensers. — With 
the exception of the 
"standard" water- 
works condenser all 
water-cooled surface 
condensers are of the 
water-tube type, that is, 
the cooling water passes 
through the tubes. Fig. 
305 shows a sectional 
elevation of the simplest 
type of surface con- 
denser. It consists es- 
sentially of a cast-iron 
shell provided with two 
heads, into which a 
number of brass tubes 
are expanded. Exhaust 
steam fills the shell and 
flows around and be- 
tween the tubes, while 
the cooling water is 
forced through the tubes 
by means of a circulat- 
ing pump. 




Dischaiige 



To Air Pump 



Fig. 305. Baragwanath Surface Condenser. 
The steam is condensed by contact with the tubes and 



drops to the bottom tube sheet from which it is exhausted by the 



524 STEAM POWER PLANT ENGINEERING 

air pump. The circulating water flows through the tubes in one di- 
rection only, hence the name ''single flow." To allow for unequal 
expansion of shell and tubes the two halves of the shell are provided 
with slightly thinner plates flanged outward, the flanges being bolted 
together with a spacing ring between them. This joint gives the shell, 
in the direction of its length, a certain amount of elasticity which is 
sufficient to allow for the greatest elongation of the tubes, without 
straining the tube sheet and causing leakage. This type of condenser 
is the least efficient of all since (1) the velocity of the water through the 
tubes is low; (2) the tubes are blanketed with a film of condensed 
steam which increases in thickness from top to bottom, and (3) air 
stagnates in the chamber opposite the air pump suction. The influence 
of these factors on the heat transmission is discussed in paragraph 242. 

Fig. 306 shows a section through a ''two-pass" condenser unit which 
is an improvement over the one just described, in that for a given tem- 
perature rise the velocity of the water through the tubes may be in- 
creased by doubling the length of its travel. In other respects, however, 
it is open to the same criticism as the single-flow device. The arrange- 
ment shown in Fig. 306 is not intended for high-vacuum work. 

Replacing the combined air and condensate pump by independent 
pumps will result in higher vacua but the tube arrangement is not 
conducive to high efficiencies. 

At the time of the introduction of the steam turbine it was discovered 
that a very high vacuum would improve turbine economies to an ex- 
tent hitherto impossible when appUed to reciprocating engines. This 
condition naturally created an era of development among the con- 
denser designers. It became evident at once that the old types that 
were capable of creating a 26-inch or 27-inch vacuum would require 
considerable modification to maintain a vacuum of 29.0 or 29.5 inches. 
Any number of condensers have been designed which are capable of 
maintaining a vacuum of 29.0 inches referred to a 30-inch barometer, 
but that the art is still in an experimental stage is evidenced by the fact 
that each new installation differs from the preceding one even for prac- 
tically identical operating conditions. Engineers are agreed that (1) 
steam should enter the condenser with the least practical resistance and 
the pressure drop through the condenser should be reduced to a mini- 
mum; (2) air should be rapidly cleared from the heat transmitting 
surfaces, collected at suitable places, freed from entrained water and 
removed at a low temperature with least expenditure of mechanical 
energy; (3) condensate should also be rapidly cleared from the heat 
transmitting surfaces, freed from air and returned to the boilers at 
the maximum practical temperature; (4) circulating water should pass 



CONDENSERS 



525 




526 



STEAM POWER PLANT ENGINEERING 



through the condenser with least friction but at a velocity consistent with 

high efficiencies. 

In the types of condensers described above the steam diminishes in 

value, due to condensation, as it passes over the tubes, hence the veloc- 
ity decreases and becomes practically 
zero at the bottom of the vessel. The 
velocity of the entrained air also de- 
creases in its passage through the con- 
denser and becomes stagnate. By shap- 
ing the condenser as shown in Fig. 307, 
the original velocity may be maintained 
to the point of air offtake. In the 
modern type of high-vacuum condenser 
the same effect has been realized by 

estabhshing steam lanes, by means of differential tube spacing or by 

a combination of both as indicated in Fig. 308. The latest practice 




Air 
Kemoval 



Fig. 



Condensate Removal 

307. Theoretically Correct 



Condenser Shape. 




Rolled Brass Tube Plate 
One R.H. 
One L.H. 
4840-1" Tubes 
2370 Upper Pass 
2160 Lower Pass 
310-H eater 



Fig. 308. Arrangement of Tubes in a Large Wheeler Surface Condenser Showing 
Steam Lane and Differential Spacing. 



CONDENSERS 



527 



is in favor of the differential spacing, that is, the tubes are spaced 
evenly across the path of the steam, leaving no preferential lanes down 
which the steam can short circuit. This uniform spacing is main- 
tained for a portion of the upper cooling surface, after which the 



Exhaust Opening 




Fig. 309. Tube Arrangement in Westinghouse Radial Flow Surface Condenser. 



distance between centers is gradually reduced. The tubes in the lower 
portion are arranged in diagonal rows in order to guide the air entrain- 
ment toward the vacuum pump suction. 

241. Cooling Water: Surface Condensers. — Since the heat absorbed by 
the cooling water must equal that given up by the exhaust, neglecting 
radiation and leakage, the amount of cooling water may be determined 
as follows: 

Hm - qi 



R = 



52 — qo 



(205) 



528 



STEAM POWER PLANT ENGINEERING 



in which 

qi, q2, and Qq = heat of liquid of the condensate, discharge and inlet 
water, respectively, B.t.u. per lb. above 32 deg. fahr. 

Other notations as in equation (203). 

Neglecting the heat content of the air entrainment and assuming a 
constant mean specific heat of unity for water, equation (203) may be 
written 



R = 



(205a) 



^2 — ^0 

in which 

ti = temperature of the condensate, deg. fahr. 

Other notations as in equation (204). 

In the ordinary low-vacuum surface condenser the depression of the 
hot-well temperature, ti, below that corresponding to the total pressure 
in the condenser may range from 10 to 25 deg. fahr. depending upon the 
amount of air entrainment and the pressure drop through the condenser. 
An average figure is 15 deg. fahr. The temperature of the discharge 
water t^ may range from 10 to 25 deg. fahr. below that corresponding 
to the total pressure in the condenser. 

The following empirical rule for determining the terminal difference 
between the temperature of the steam corresponding to the vacuum 
in the condenser and that of the circulating water discharge gives 
results agreeing substantially with current surface condenser practice 

^ td = t- to, (206) 

in which 

td = terminal difference, deg. fahr., 
t = temperature corresponding to saturated vapor pressure 

(Po + B), 
to = initial temperature of the circulating water, deg. fahr., 
po = pressure of saturated vapor corresponding to temperature ^o, 
B = coefficient, as follows: 

VALUE OF COEFFICIENT B. 



Vacuum, In. 


B. 


Vacuum, In. 


B. 


Vacuum, In. 


B. 


1.00 
1.25 
1.50 


0.20 
0.25 
0.30 


1.75 
2.00 
2.50 


0.35 
0.40 
0.45 


3.00 
3.50 
4.00 


0.50 

0.60. 

0.70 



Thus for ^0 = 70 and a 2-inch vacuum: p = 0.739,^5 = 0.40 t cor- 
responding to 0.739 + 0.40 (= 1.139) = 83.0 deg. fahr., whence td = 83 
- 70 = 13 deg. fahr. 



CONDENSERS 529 

Example 45. (Low-vacuum condenser.) Required the weight of 
coohng water necessary to cool and condense one pound of steam under 
the following conditions: Engine uses 16 lb. steam per brake hp-hr., 
initial pressure 140 lb. per sq. in. absolute, quality 0.99, initial tempera- 
ture of the cooling water 70 deg. fahr., vacuum 26 inches referred to a 
30-inch barometer. 

From the Mollier diagram or by calculation from steam tables Hi = 
1185 (approx.), ts= 126, Hr by assumption = 1 per cent of //». 

From equation (145) 

2^46 
^ = 1185 - 0.01 X 1185 - ^^ = 1014. 

16 

Assume 

^, = ^^ - 15 = 111, ^2 = ^3 - 20 = 106 [see equation (206)]. 
„ 1014- 111 H- 32 _._,, 
^= 106-70 =^^'^^^' 

With the modern high-vacuum surface condenser in connection 
with a practically air-tight system the temperature of the condensate 
will be from to 5 degrees lower than that corresponding to saturated 
vapor at condenser pressure and the temperature of the discharge water 
will range from 2 to 10 degrees below that corresponding to the vacuum. 
The pressure drop through the condenser from exhaust inlet to air 
pump suction varies with the type and size of condenser and the rate of 
driving and ranges from 0.02 to 0.2 inch with an average at rated 
load of approximately 0.1 inch. 

Example 46. (High-vacuum surface condenser.) Required the 
weight of cooling water necessary to cool and condense one pound of 
steam under the following conditions: Turbine uses 12 lb. steam 
per kw-hr., initial pressure 200 lb. per sq. in. absolute, superheat 150 
deg. fahr., initial temperature of cooling water 70 deg. fahr., vacuum 
28.5 inches referred to a 30-inch barometer. 

From steam tables, Hi = 1283. Assume Hr = 0.5 per cent of Hi. 

Q412 
From equation 145, H = 1283 - 0.005 (1283) - ^^ = 993. 

Assuming a pressure drop of 0.1 inch, the probable absolute pressure 
in the condenser will be 30 — (28.5- +0.1) = 1.4 in. The correspond- 
ing temperature of vapor at this pressure ts = 89.5 deg. fahr. ^>>-t-' 

Assume ^^ = ^^ - 4 = 85.5, t2 = t^ - S = 81.5. ^ . , ^ r^^ -' 1) ^ "^ 

r. 993-85.5+32 oi-t,u '^ 

Whence R = ^j-^ =^r = 81.7 lb. 

oi .0 < U 

243. Heat Transmission througli Condenser Tubes. — Numerous in- 
vestigations have been conducted on special laboratory apparatus and 
on condensers in actual service for determining the heat transmission 
through condenser tubes, but the laws based on these results have been 
far from harmonious. Jn steam engine practice where the vacua are 
comparatively low extreme refinement in design is unnecessary and 



530 STEAM POWER PLANT ENGINEERING 

simple empirical formulas for estimating the extent of cooling surface 
are sufficiently accurate. In modern high-vacuum practice, however, 
particularly for large turbo-generators where a fraction of an inch of 
change in vacuum greatly affects the economy of the prime mover, and 
where thousands of square feet of cooling surface are involved in a 
single unit the older empirical rules are apt to lead to serious error. 
Despite the tremendous advance in condenser design during the past 
few years the art is still largely a matter of experience and the best 
rules are subject to arbitrary assumptions. 

In any type of surface condenser, neglecting radiation and leakage, 
the heat absorbed by the cooling water, SUd, must be equal to that 
given up by the exhaust Wm, {H^ — qi) or, 

in which SUd = Wm {Hm - 5i),* (207) 

S = extent of cooling surface, sq. ft., 

U = experimentally determined mean coefficient of heat trans- 
mission, B.t.u. per hour, per deg. fahr. difference in tem- 
perature, d, per sq. ft., 

d = mean temperature difference between that of the steam and 
of the circulating water, deg. fahr., 
w^ = weight of condensate, lb. per hr. plus the air entrainment, 
Hm = heat content of the exhaust steam, moisture and air entrain- 
ment, B.t.u. per lb. above 32 deg. fahr., 

^1 = heat of hquid of the condensate. 

From equation (207) .S = " ^^^^^'^'^ (208) 

In view of the liberal factor allowed in estimating the value of U 
and because of the uncertainty of the true value of d, the influence of 
the heat content of the air entrainment becomes negligible and equation 
may be written: 

S = Hifl^A±i?), (209) 

in which ^^ 

w = weight of condensate, lb. per hr., 

H = heat content of the exhaust steam, B.t.u. above 32 deg. fahr., 
ti = temperature of the condensate, deg. fahr. 

Since the heat absorbed by the cooling water is equal to that given 
up by the steam, equation (209) may also be stated 

o Q {k — to) /oim 

S ^j^—, (210) 

* This is on the assumption that the heat transfer is directly proportional to the 
mean temperature difference. See also equation (223). 



CONDENSERS 



531 



in which 
Q = total weight of cooling water, lb. per hour, 
^2 = temperature of the discharge water, deg. fahr., 
to = temperature of the intake water, deg. fahr. 

Considering first the coefficient of heat transfer, it must be remembered 
that the coefficient U, as used in equations 207-210, refers to the mean 
or average value for the entire surface and not the actual value, since 
the latter varies widely for different parts of the condenser. The actual 
value varies from more than 1000 for air-free vapor, in the first few 
rows of tubes (where the steam comes directly into contact mth the 
cooling surface) to less than 50 in the bottom row (where the tubes 
may be practically blanketed with the condensed steam) and to 3 or 
less for tubes surrounded only by air. Tests by various investigators 
show that the actual value of U for a given temperature difference varies 
with 

(a) material, thickness, size, shape and cleanliness of the tubes; 

(h) velocity of water through the tubes; 

(c) percentage of air on the steam side of the tubes; 

(d) critical velocity of the water in the tubes; 

(e) extent of water blanketing on the steam side of the tubes; 
(/) viscosity of the circulating water. 

Taking the material coefficient, m, of plain copper tubes as 1.00, under 
similar conditions the heat transfer for other materials is approximately 



.Ft. per Ilr. 

i i 


\, 






DECREASE IN HEAT TRANSMISSION 
OF 




^ 




SERVICE IN CONDENSER. ALL ADMIRALTY 
MIXTURE TUBES 


B.t.u.Transferred perSq 
per Deg. Dlf. = 








^\ 




^v. 


















\ 


^ 


^s. 


















\ 


\ 














N.E.L'A.May 22, 1916 



2000 4000 6000 8000 

Number of Hours in Service in Condenser 

Fig. 310. 



as follows: Admiralty brass 0.98, Muntz metal 0.95, tin 0.79, Admi- 
ralty lead line 0.79, Monel metal 0.74 and Shelby steel 0.63. No better 
material than Admiralty brass has been found and it is the standard for 
modern condenser practice. Corrosion, oxidation and pitting have a 



532 



STEAM POWER PLANT ENGINEERING 



marked effect in reducing the heat transference and may lower the 
conductivity as much as 50 per cent. (See Fig. 310.) The cleanhness 
coefficient, c, is about 0.9 for such waters as New York or Chicago. 
The coefficient of heat transfer appears to decrease with the in- 
crease in diameter but since the one-inch tube, No. 18 B.W.G. is 
the most common in use this factor need not be considered for any 
other size. 

The influence of the velocity of the water through the tubes on the 
coefficient of heat transfer is illustrated in Fig. 311 and Fig. 315. Ac- 
cording to Orrok the value 
of Uj other conditions re- 
maining constant, varies ap- 
proximately as the square 
root or six-tenths power of 
the velocity.* For the ordi- 
nary low-vacuum condenser 
the velocity through one- 
inch standard tubes seldom 
exceeds 3 ft. per second, 
whereas velocities as high 
as 10 ft. per sec. are not 
uncommon in the high- 
vacuum type. An average 
value for the latter is 8 ft. 
per sec. Except for a very 
low rate of flow (below that 
in average condenser prac- 
tice) critical velocities need 
not be considered. For ex- 
ample, critical velocities for 
a one-inch No. 18 B.W.G. condenser tube are approximately as foflows.* 































10. 
















A 


J^ 


^2 




3 




y 


r 


^700 












/ 


y 


X 
^ 






^ 


y 














A 


/- 


y 






y 










a 600 








/. 


'/> 


/ 




^ 


y 




.5 






> 






A 


^/ 


/ 




/ 




^ 




^ 


^ 






-a 

|500 

C 




j 


y.\ 


/ 








€^ 


^^ 




^ 


-^i 








1 


/ 




/^ 




















tm 

d 






/ 


y 


^ 




^ 


^8 














/ 


1 


4 


Y 






^ 
















i 

.300 


I 


A 


y/ 


/ 


/ 






















'//h 


/ 


/ 








1 S» 

2 J 

3 W 

4 H 


r 




a 


3sse_ 
/■ei'cl 




g 




h 


/ 










ton 




2 200 


epburn 




1 












5 Hage 

6 Stantc 


■n !in n 






















7 .rm.l« 

8 Allen 


















9 Clement & Garland 

10 Orrok 

































Fig. 311. 



L 2 3 4 5 6 7 

Velocity circulating water-ft.per sec. 

Variation of Heat Transmission with 
Water Velocity. 



ira. 


. 40 


50 


60 


70 


80 


90 


100 


115 


120 


150 


Vc.. 


..0.50 


0.42 


0.36 


0.32 


0.28 


0.25 


0.22 


0.19 


0.17 


0.14 


Va. 


..2.84 


2.40 


2.06 


1.81 


1.58 


1.42 


1.27 


1.09 


0.94 


0.80 



in which 

tm = mean temperature of the water, deg. fahr., 

Vc = the lower critical velocity, ft. per sec, below which all motion 

is stream-line unless disturbed artificially, 
Va = the high critical velocity above which all motion is turbulent. 

* The proportioning of Surface Condensers, Geo. A. Orrok, Journal of A.S.M.E,, 
Nov., 1916. 



CONDENSERS 



533 



Heat Transference for Air 



The effect of air on the heat transference is very marked as is shown 
in Fig. 314. The depression of the hot-well temperature below that 
corresponding to the vacuum 
may be reduced by good design. 
Certain designs of dry tube con- 
densers may give hot-well tem- 
peratures somewhat higher than 
the average temperature in the 
condenser and tests have been 
reported on several other de- 
signs in which the depression 
was zero. Orrok's investigations 
show that air entrainment re- 
duces the heat transference ap- 
proximately according to the 
law {pv -^ pcY, in which p^ = ^^^- ^^^ 
pressure of the vapor and pc the total pressure in th( 




30 



25 20 15 10 5 Atm. 

Air Pressure*- Inches of Mercury 

Heat Transmission Steam to Air. 



The value of (p 

neat Transference for Air 




10 20 30 40 50 60 
liean Rate of Flow of Air ^Feet per Second 



condenser. 
Pc^ varies within wide limits, but for tight con- 
densers with efficient air pumps 
it may be taken as 0.95. 

The reduction in heat trans- 
mission due to thicloiess of water 
film on l)oth sides of the tubes 
has been expressed mathemati- 
cally but it is customary in con- 
denser design to include this 
factor in the assumed value 
of U* 

The coefficient of heat trans- 
mission increases with the mean 
temperature of the circulating 
water, that is, the warmer the 
water and the lower the vacuum 
the smaller will be the mean 
temperature head required to 



70 



Fig. 313. Heat Transmission Steam to Air. 

transmit practically constant amount of heat through the surface. 



According to Orrok 



U = kcpm 



,0.6 



dk 



(211) 



♦Trans. A.S.M.E., Vol. 35, 1915, p. 67. 



534 



STEAM POWER PLANT ENGINEERING 



ffi 500 



1 




1 




1 




\ 




± 




^ it 




"^ 




js 




^^ 




-l^-U 




^^ 




5». 




^^ 








^S^*^:^ 




^^r^ 


^^:=:- - 




— - ~~= — ^^S- 











5 10 15 20 25 30 35 40 45 50 

Depression of Hot Well Water Temperature below Vacuum Temperature 
Deg. Fahr. 

Fig. 314. Relation Between Coefficient of Heat Transfer and Temperature 

Depression. 




2 3 4. 5 6 7 8 9 1 15 

Mean Velocity of Circulating. )J7ater, 

Feet per Second 

Fig. 315. Rate of Heat Transfer Versus Circulation Water Velocity, 
of Tests by Geo. A. Orrok, Trans. A.S.M.E., 1910. 



Results 



CONDENSERS 



535 



in which 

U = mean coefficient as previously defined and as used in con- 
nection with equations (209) and (210), 
A; = experimentally determined coefficient = 350 for average 

working conditions, 
c = cleanliness coefficient, 
p = air richness ratio = (p„ ^ pj^, 
m = material coefficient, 
V = velocity through the tube, ft. per sec, 
d = logarithmic mean temperature difference. 

The following empirical rule gives values of U which agree sub- 
stantially with current practice in condenser design 

U = KVv (212) 

VALUE OF K FOR VARIOUS INITIAL TEMPERATURES OF CIRCULATING WATER. 



Initial Temp., 
Deg. Fahr. 


K. 


Initial Temp., 
Deg. Fahr. 


K. 


Initial Temp., 
Deg. Fahr. 


A'. 


40 
45 
50 


141 
160 
175 


60 

65 
70 


192 
198 
203 


80 
90 

100 


212 
218 
220 



Mean Temperature Difference. — It is definitely known that the 
quantity of heat passing through the cooling surface is proportional to 
some power of the temperature difference at any instant, but the in- 
stantaneous temperature difference is indeterminate, consequently it is 
necessary to establish an average or mean temperature difference for 
the whole period of thermal contact of the steam and circulating water. 

If ts = temperature of the steam or hot substance, 

t = any momentary temperature of the circulating water, 
^0 = initial temperature of the circulating water, 
<2 = final temperature of the circulating water, 
d = mean temperature difference, 
Q = weight of circulating water, lb. per hr., 
S = extent of cooling surface, sq. ft. 

Ui = instantaneous value of the coefficient of heat transfer, 
U = mean coefficient of heat transfer for the entire period of heat 
exchange. 

All temperatures deg. fahr. 

Then the heat transmitted per hour through the elementary surface 

dS is C7i (t, - t). 



536 



STEAM POWER PLANT ENGINEERING 



Since the temperature rise for this period is dt the heat absorbed by 
the circulating water per hour is Q dt (theoretically this should be cQ dt 
in which c is the mean specific heat of the water, but for all practical 
purposes the value of c may be taken as unity). 



Cubic Feet Free Air per Minute-N.Y. Edison -Ci 
1 2 3 4 5 e 




Cubic Feet Free Air per Minute-Detroit Edison-Curves 1- S-'^land 5 

Fig. 316. Curves Showing Effect of Air Leakage on Condenser Efficiency, 



These two quantities must be equal, or 

Ui{ts-t)dS = Qdt. 



from which 



dS = 



Q dt 



(213) 

(214) 



U, ts - t 

If the temperature of the steam is assumed to be constant ts is inde- 
pendent of t, and if the heat transmitted per hour is assumed to be 
directly proportional to temperature difference, U is likewise independ- 
ent of t and Ui = U, therefore the relation between rise in temperature 
of the circulating water and the surface traversed becomes 

ts — to 



Q , 

For the whole period of transfer, 

SUd = Q{t2- to), 
Q {ti - to) 



d = 



US 



Combining equations (216) and (218) and reducing, 

7 t'y — to 



lOge 






(216) 

(217) 

(218) 

(219) 



CONDENSERS 



537 



This is known as the logarithmic mean temperature difference and is 
the one most commonly used in condenser design. The relation be- 
tween temperature of the steam and that of the circulating water for 
this condition is shown graphically in Fig. 317. 

















Stcaiu in Condenser 
















u 










































J3 










































.^=i^ 


^ 
































__^. 


-- 


-;: 








o 






















.J 


,^' 


'^Ltc- 


^ 














p 














^^ 


^- 


-' 


'c\t 


cu\a 


tAng 























^ 


f 


,-■ 


■^ 


■' 




— 

























\^ 




a 








































H 




*0 












































JL. 




_^ 





































Length of Water Pass Through Condenser 

Fig. 317. Rise of Circulating Water Temperature in Condenser Tubes. 

If the rise in temperature of the circulating water follows the dotted 
line cd the mean temperature difference may be expressed 



d = L- 



h + k 



(220) 



This is known as the arithmetic mean temperature difference and is 
used only for rough calculation or where other influencing factors can 
only be approximated. 

If the quantity of heat transmitted per hour is proportional to the 
nth power of the instantaneous temperature difference, as appears to 
be the case in actual practice, and U is assumed to be constant 

Qdt= U{t,-t)-dS (221) 

Integrating and reducing, 



By assumption, 



s = % Kt, - toY-^ - iu - t,y-^] -^> 



Therefore 



SUd- = Q(t2- to). 
d = \- ^^ ~ "") ^^' ~ ^«) 



w^-" - its - t^y 



JS 



(222) 
(223) 
(224) 



This is known as the exponential mean temperature difference. Or- 
rok's * experiments lead to a value of n = 0.875. Loeb f assigns a 
value oi n = 0.9. Because of the uncertainty of the value of U it is 
sufficiently accurate for most purposes to take n = unity, which results 
in the logarithmic mean temperature difference. 

* Jour. A.S.M.E., Nov., 1916. f Jour. Am. Soc, Naval Engrs., Vol. 27, May, 1915. 



538 STEAM POWER PLANT ENGINEERING 

Orrok * gives a general rule for high-vacuum surface condensers 
operating under favorable conditions which may be reduced to the form 

S = 4of^ [its - to)i - {t, - fe)i] . (225) 

Let d = outside diameter of the tube, in., 

n = number of tubes in each pass of the condenser. 
I = length of water travel, or total tube length, ft. 

Then, S = "^ nl, whence I = 3.83 S -^ dn. (226) 

By simple arithmetical calculation it may be shown that 

""^ 123^v{d-2ty' (^^'^^ 

in which t = thickness of the tube, inches. 

Example 47. (Low-vacuum condenser.) Approximate the amount 
of cooling surface for a 1000-hp. compound engine operating under 
the following conditions: Water rate 16 lb. per hp-hr., initial steam 
pressure 140 lb. absolute, initial quaUty 0.99, inlet temperature of 
circulating water 70 deg. fahr., vacuum 26 in. referred to a 30-inch 
barometer. 

In view of the absence of data, in this particular problem, for esti- 
mating the value of U with any degree of accuracy it is sufficiently 
accurate to assume the temperature of the steam in the condenser to 
be that of saturated vapor corresponding to the vacuum, and for the 
same reason the heat of the exhaust may be assumed to be that of 
saturated steam corresponding to the absolute pressure in the condenser. 

The temperature of vapor ts corresponding to an absolute pressure of 
4 inches (p^ = 30 - 26) is 126 deg. fahr. and i^ = 1114 (approx.). 

In the ordinary engine condenser considerable air will be carried with 
the steam into the condenser and the hot-well depression may range 
from 5 to 20 degrees; assume the depression to be 10 degrees, then 
t^ = ts - 10 = 126 - 10 = 116 deg. fahr. 

Any value may be assumed for ^2 greater than ^0 and less than ti. 
The nearer ^2 is to to the greater must be the quantity of circulating 
water per lb. of condensate. On the other hand, the nearer ^2 is to ^0 
the less is the mean temperature difference d, and hence the greater 
must be the cooling surface for a given weight of condensate. When 
water is cheap and the head pumped against is small ^2 may be given a 
lower value than when water is costly and the discharge head is large. 
In average engine condenser practice ^ may range from 5 to 20 degrees 
below ^1; assume it to be 10 degrees, then ^2 = ^1 — 10 = 116 — 10 
= 106 deg. fahr. Equation (206) gives ^2 = 105.5. 

Because of the great latitude in assuming values of ^1 and t^ it is suffi- 
ciently accurate to use the arithmetical mean, or 

,= 126 -Z0±106^ 38.0. 

z 

* Jour. A.S.M.E., Nov., 1916. 



CONDENSERS 539 

In engine practice a very liberal factor is allowed in assuming a value 
for U because of the possible reduction in heat transmission caused by 
the deposit of cyhnder oil on the tubes and because of the air entrain- 
ment. For the usual engine type of condenser a safe value is [/ = 300. 
According to equation (212), U = 300 for v = 2.25 ft. per sec. 

Substituting these values in equation (209) and reducing 

16,000(1114-116 + 32) 
S = — 3^^^^^g = 1446 sq. ft. 

This corresponds to approximately 11.0 lb. of condensate per hr. 
per sq. ft. of tube surface. An average figure commonly quoted for en- 
gine condensers is 10 lb. of steam per hr. per sq. ft. of tube surface for 
24-26 in. vacuum with 70-degree cooling water. A rough rule is to 
allow 2 sq. ft. of cooling surface per i.hp. 

Example 48. (High- vacuum condenser.) Calculate the amount of 
tube surface required for a 10,000-kw. turbine operating under the fol- 
lowing conditions: Water rate 12.0 lb. per kw-hr., initial absolute 
pressure 200 lb. per sq. in., superheat 150 deg. fahr., temperature of 
circulating water 70 deg. fahr., vacuum 28.5 inches referred to a 30- 
inch barometer, water velocity through tubes 8 ft. per sec. Cooling 
surface to consist of one-inch (18 B.W.G.) Admiralty tubes. 

For maximum theoretical efficiency t2 = h = ts. This condition 
i^ possible only for air-free vapor, perfect heat transmission, and no 
pressure drop between turbine nozzle and air pump suction. In the 
very latest designs the pressure drop between turbine nozzle and air 
pump suction seldom exceeds 0.2 in. The temperature of the con- 
densate varies from ^i = ^^ — to ^i = ^s — 4 deg. fahr., and (2 varies 
from ^2 = ^s — 4 to ^s — 10 deg. fahr. Assume a pressure drop of 0.2 
in., ti = ts and (2 = ts — 8, then 

Ps = 30.0 - (28.5 + 0.2) = 1.3 in. and the corresponding ts = 87.1 
deg. fahr. 

For the given conditions H = 993 (see example 46). 

Then Q = 12x10,000(993^-87.1+32) ^ ^^^^^^ ,^_ 

'°^'87.1-79.1 

Exponential mean gives d = 11.97 deg. 

The condenser must be designed for the maximum load when the 
circulating water is at its highest temperature and a suitable factor 
should be allowed for dirty, oxidized tubes and the presence of undue 
amounts of air. For this reason a much lower value of U is assumed 
than is possible with everything in first-class shape. An average 
value for a velocity of 8 ft. per sec. is U = 600. According to equation 
(212) U = 575. 

Substituting these values in equation (210), 

„ 12,368,000(79.1 -70) ^ _ „_ ,^ 
^ = 600 X 11.98 = ^^'^^^^^- ^'' 



540 



STEAM POWER PLANT ENGINEERING 



Corresponding to 1.56 sq. ft. per kw. of turbine rating. Surface 
condensers for large turbines have generally from 1.6 to 1.7 sq. ft. of 
condensing surface per kw. See Table 93. 

Orrok's rule gives for this example 



S = 



12,368,000 



[(87.1 - 70)« = (87.1 - 79.1)«] 



40.3 X 80.6 
= 11,500 sq.ft. 

Taking the heat content of the steam as that of saturated steam 
at condenser pressure Orrok's rule gives S = 12,650 sq. ft. In fact, 
Orrok's rule is based on the assumption that the steam entering the 
condenser is saturated, an assumption which simplifies calculation and 
which is justifiable in view of the uncertainty of the true values of many 
of the factors entering into the problem. 



TABLE 92. 

TEST OF 50,000 SQ. FT. SURFACE CONDENSER, 74TH ST. STATION 
INTERBOROUGH RAPID TRANSIT CO. 

(H. G. Stott and W. S. Finlay, Jr.) 

Pressure at throttle, lb. abs 220 

Temperature at throttle, deg. fahr 487 

Superheat 97 

Load, average kw 31,233 

Exhaust vacuum, in 28. 61 

Exhaust pressure, in. abs 1 . 39 

Corresponding temp. deg. fahr 89.4 

Mean temp. diff. (log.) 12. 9 

Condensate, lb. per hr *. . 357,060 

B.t.u. per sq. ft. per hr. per deg. mean temperature difference. . . 490 

Air leakage, cu. ft. per min 16. 88 

Temp, of hot well, deg. fahr 86 

Temp, intake water, deg. fahr 70. 8 

Temp, discharge water, deg. fahr 80 . 9 

Temp, rise, deg. fahr 10 . 1 

Circulating water, gal. per min 64,700 

Ratio circulating water to condensate 91 

TABLE 93. 
MODERN SURFACE CONDENSER PROPORTIONS. 



Size of Turbo- 
generator. 


Tube Surface, 
Sq. Ft. 


Sq. Ft. Tube 
Surface Per Kw. 


Size of Turbo- 
generator. 


Tube Surface, 
Sq. Ft. 


Sq. Ft. Tube 
Surface Per Kw. 


500 
1000 
2000 
5000 


1,500 

2,750 

5,000 

10,000 


3.00-3.50 
2.75-3.25 
2.50-3.00 
2.00-2.50 


10,000 
15,000 
20,000 
35,000 


17,500 
25,000 
32,000 
56,000 


1.75-2.25 
1.67-2.00 
1.60 
1.60 



The curves in Fig. 318 are based upon equation (209) with U = 300 
and afford a simple means for determining the extent of cooling surface 



CONDENSERS 



541 



for different conditions of operation. For any other value of U mul- 
tiply by 300 and divide by the new value of U. 

Design and Proformance of Surface Condensers: Jour. A.S.M.E., Nov., 1916, p. 864; 
Power, Aug. 29, 1916, p. 300; Jour. A.S.M.E., Jan., 1916, p. 23; Jan., 1915, p. 546; 
Aug., 1915, p. 459. 

Surface Condenser Air Pumps. — See paragraphs 308 to 310. 



ajoiQ 



0.018 



Q.017 



0.016 



D.Q15 



§ 0.014 

o 

t; 0.013 

S 0.012 
^ Q.01X 



o 0.010 

§5 0.009 

© 
o 



r*^ 0.006 

<^ 0.005 
0.004 
0.003 
0.002 
0.001 
0,000. 































y 


y 


























/ 


/ 


y^ 

y 






















\y 


t\ 






/ 


y 


y 














rN. 


1 


f 








/ 


y' 


y 


/ 










^ 


.^ 


> 








> 


/ 




^^ 


'-' / 


c 








.^ 


r 








y 


/ 








y 


y y 


y 




A 


y 






/ 


/ 


/ 








y 


/ 




yy. 


(y 


f> 


>f 




/ 


/ 


y 






/ 


y 


y 


/ 


/ 


^ y 
y ^ 


<yy 


^ 






/ 


y 






y 


/ 


/ 


/ 


y 


/ 


/ 


y>. 


y<y 


y 










/ 


/ 




y 


/ 


/ 


>( 




/ 


y<^ 


^yy 

'y y 


y 






y 


y 




/ 


/ 


/ 


y 

y 


/" 


Xo 

^ 




/ 


<y 


y 

v^' y 


y 




/ 


/ 


/ 


> 


y 


/ 


y 


/ 


'y. 


% 


:^ 


^ 


y 


' yy 


J; 


y 


y 


/ 


/ 


') 




/ 


/ 


0. 


^ 


'4 




y 


y .^ 


y-Oy 




^ . 


y \ 


/: 




^ 


y 

/ 




y 


y 




f 


^ 


/ 


<>: 


y yy 

'y'y<y 


/. 


^y 


'/, 


g 


^ 


^ 


y 




y 


/ 


P 


y 


/ 


yy. 


^yyy 


y ^ 


X> 


/ y 






/ 


y 


/ 


/ 


y 


y 




/ 


/ 


yy 


y^^ 


^ 


y/^ 




^ 


V 


y 


y 


/ 




< 




1 






^:>^ 


'yy^ 


yy 


O^y!/" 




/ 


y 
y 


/ 


y 


y 


/ 


^ 


^ 


% 




/ 


^/4 

^y'y/ 


^^^P 


r 




/ 




y 
^ 




y^ 


y 


^ 


y 


1 


fe 


^y 


m 


^^^^ 



90. 
100, 

not 

120' 
130 
150 
170 
200' 

250" 



10" 15° 20' 25' 30° 35 40 4-550 60" 70 80 90100 120140100180200^ 

Initial Temperauure difference between Steam and Water 

Fig. 318. Curves for Determining the Amount of Cooling Surface. 

243. Dry-air Surface Condensers (Forced Circulation). — Where water 
is very scarce and the feed supply is reclaimed by condensing the ex- 
haust steam, water-cooled condensers may be prohibitive in cost of 
operation, even when combined with cooling tower or other water-cool- 
ing device, since the latter involves a loss of water approximately 
equivalent to the amount of steam condensed, due to evaporation. 

A notable installation of air-cooled surface condenser is that in an 
electric station of 2000-horsepower capacity in the city of Kalgoorlie, 



542 



STEAM POWER PLANT ENGINEERING 



West Australia.* The condenser consists of a large number of narrow 
chambers constructed of thin corrugated sheet-steel plates spaced J 
inch between centers. Each chamber has 1345 square inches of coohng 
surface. Fifty-one of these chambers are grouped into a compartment 
and 15 compartments constitute a section. Each section is equipped 
with three motor-driven fans 7 feet in diameter and running normally 
at 320 r.p.m. In all there are six sections, giving a total cooling sur- 
face of 45,000 square feet. The steam consumption of the main engines 
is 16 to 16.5 pounds per i.hp-hour at rated load. At full load the fans 
require 130 kilowatts, or approximately 10 per cent of the station out- 
put. The average vacuum obtained is about 18 inches throughout 
the year and ranges from inches on very hot days to 22 inches in cooler 
weather. The following figures, based on actual observation, show the 
effect of temperature of the external air on the vacuum when condensing 
32,000 pounds of steam per hour (the rated capacity of the condenser). 



Temi)erature Ex- 


Vacuum, Inches 


Temperature Ex- 


Vacuum, Inches 


ternal Air, 


(referred to 30-Inch 


ternal Air, 


(referred to 30-Inch 


Degrees F. 


Barometer). 


Degrees F. 


Barometer). 


42.8 


22 


96.8 


9.6 


50 


21.2 


100.4 


7.6 


60.8 


20 


107.6 


3.6 


68 


18.4 


113 





78.8 


16 







Air-Cooled Surface Condensers: Engineering News, Oct., 1902, p. 271; ibid., Vol. 
49, p. 203. 

244. Quantity of Air for Cooling ^Dry-air Condenser). — The volume 
of air, under atmospheric conditions, necessary to condense steam to 
any given temperature may be determined as follows: 

Let H = heat content of the steam at condenser pressure, 
ts = temperature of the vapor in the condenser, 
^1 = temperature of the condensed steam, 
t = temperature of the air entering condenser, 
to = temperature of the air leaving condenser, 
V = volume of air in cubic feet necessary to condense and cool 

one pound of steam, 
B = specific weight of air under atmospheric conditions, 
C = mean specific heat of air under atmospheric conditions, 
d = mean temperature difference between the air and steam, 
S = cooling surface in square feet, 

U = coefficient of heat transmission, B.t.u. per square foot per 
degree difference in temperature per hour. 

* This condenser has been recently discarded since the cost of water has been 
greatly reduced, 



I 



CONDENSERS 543 

Since the heat absorbed by the air must be equal to the heat given 
up by the steam, neglecting radiation, we have 

VBC (fo-t) = H -h + 32, (228) 



from which 



^-"im^- ^'"'^ 



For practical purposes C may be taken as the specific heat of dry air, 
the error due to this assumption being negligible even if the air is satu- 
rated with moisture. 

Example 49. How many cubic feet of air are necessary to condense 
and cool one pound of steam under the following conditions: Vacuum 
20 inches; temperature of entering air, leaving air, and condensed steam, 
60, 110, and 140 deg. fahr., respectively? 

Here H = 1130 (from steam tables), 

^0 = 110, h = 140, ^ = 60, C = 0.24, B = 0.075. 

Substituting these values in equation (229), 

^ = 0.07"x°oT24aiO-60) ^ ^^^^ ""''''' ^"^^ °^ "-'' necessarj' to 
condense one pound of steam under the given conditions. 

The proper area of cooling surface depends upon the value of the 
coefficient of heat transmission, which varies with the velocity and 
humidity of the air and character of the coofing surface. Accurate 
data are not available on this point. 

A few experiments made at the Armour Institute of Technology 
gave values oi U = 10 to 25 B.t.u. per hour, per square foot, per degree 
difference in temperature for air velocities of 500 to 4000 feet per minute, 
for corrugated-steel sheeting J inch thick. Assuming these values of U 
for the above example, >S = 1.5 square feet of cooling surface per pound 
of steam condensed per hour for air velocity of 4000 feet per minute, 
and *S = 3.7 square feet for a velocity of 500 feet per minute. 

345. Saturated-air Surface Condensers (Natural Draft). — Fig. 319 shows 
vertical and horizontal sections of a Fennel saturated-air surface con- 
denser. The apparatus consists of an upright cylindrical shell contain- 
ing a number of vertical 4-inch steel tubes through which air is drawn 
by natural draft. A centrifugal pump circulates about one half gallon 
of water per horsepower per minute from a cistern below the condenser. 
The water flowing over the upper tube sheet and then descending the 
tubes by gravity forms a film over their entire interior surface. 

The condensing action is as follows: The current of exhaust steam 
entering the side of the shell at A is caused by suitable baffle plates to 
circulate among the tubes, and in condensing gives up its latent heat 
to the water film, which wholly or partially evaporates, saturating the 
ascending current of air at its own temperature. The upward current 



544 



STEAM POWER PLANT ENGINEERING 



of hot vapor-laden air carries off tlie heat into the atmosphere. The 
cooling water which is not evaporated and lost to the atmosphere falls 
into the cistern below to be again taken up by the circulating pump, 
the water level in the cistern being kept constant by a float governing 




- 1— - ^ 





ill 




\ 
\ 



Fig. 319. Fennel Saturated-air Surface Condenser. 



a valve on the supply pipe. The non-condensable gases collect at C, 
where they are removed by the dry-air pump, while the condensed steam 
is drawn off from the bottom tube sheet by the vacuum pump and 
discharged into the hot well. An excellent feature of this device is 

that the film of water on the 
cooling surface is secured with- 
out interference with the ascend- 
ing air currents and also without 
the use of sprays through small 
orfices likely to become clogged 
with rust or sediment. Where 
the recovery of the condensed 
steam is essential and a high 
vacuum of secondary importance, 
condensers of this type have 
proved to be good investments 
on account of the low first cost. 
Table 94 gives the results of a 
test of a condenser of this type, 
taking steam from a 30-in. by 58-in. by 48-in. engine running at 45 r.p.m. 
(Power, December, 1903, p. 672; West. Elect., May 19, 1900, p. 323.) 
Fig. 320 illustrates the Pennel ''flask" type of atmospheric con- 
denser. The exhaust steam enters below and follows the zigzag course 




Fig. 320. 



Pennel Flask Type of Saturated- 
air Surface Condenser. 



I 



CONDENSERS 545 

bounded by the internal stay channels, condensing as it goes and driving 
before it the non-condensable gases to the outlet at the top. The con- 
densed steam gravitates to the bottom and thence to the hot well. The 
top of the flask is trough shaped and causes the cooling water to flow 
down the sides of the flask in a thin stream. The portion of the cool- 
ing water not evaporated collects at the bottom of the flask and flows 
to the cooling-water reservoir. 

TABLE 94. 
TEST OF FENNEL SATURATED-AIR SURFACE CONDENSER. 

Duration of trial 9 hours 

Average steam pressure at engine by gauge 139 . 8 pounds 

Average vacuum, mercury column 17.5 inches 

Average temperature in condenser 123.7 deg. fahr. 

Average temperature of circulating water 116.4 deg. fahr. 

Average temperature of city water 52 deg. fahr. 

Average temperature of outside air 62 deg. fahr. 

Average temperature of saturated air 106 deg. fahr. 

Average draft in stack of condenser 1.1 inches 

Average humidity of outside air 67 per cent 

Average amount of steam condensed per hour 7950 pounds 

Average amount of circulating water used per hour 114,660 pounds 

Average amount of city water used per hour 3462 pounds 

Pounds of steam per pound of city water 2.3 

Pounds of circulating water per pound of steam 14. 4 

Average horsepower of engine 569 . 7 

Steam, pounds per i.hp-hr 13 . 95 

Horsepower required to run air pumps 10 . 5 

Horsepower required to run circulating pumps 3.0 

Condensing surface, square feet 3900 

Pounds of steam condensed per square foot surface per hour 2038 

Barometer 28 . 58 inches 

Vapor tension corresponding to 123.7 degrees 3.82 inches 

Per cent of main engine steam used by auxiliaries 2 . 38 

246. Evaporative Surface Condensers. — An evaporative surface con- 
denser consists of a number of copper, brass, wrought- or cast-iron 
tubes arranged horizontally or vertically and connected to manifolds 
or chambers at each end. The exhaust steam passes through the 
tubes and a thin film of water is allowed to flow over the external sur- 
faces. The cooling effect is brought about by the evaporation of 
part of the circulating water, and the general principle of operation is 
the same as that of the saturated-air condenser described above. Evapo- 
ration is sometimes hastened by constructing a flue over the tubes, 
thereby creating a natural draft, or by means of fans. With hori- 
zontal cast-iron tubes and natural draft, vacua from 23 to 27 inches 
are readily maintained with a cooling surface of approximately eight 



546 



STEAM POWER PLANT ENGINEERING 



tenths square foot per pound of steam condensed per hour. With 
vertical brass tubes and fan draft 8 pounds of steam per hour per 
square foot of coohng surface is not an unusual figure. The amount 
of cooling water evaporated per pound of steam varies from eight 
tenths to one pound, depending upon the draft. The power necessary 
to operate the pumps and fans varies from 1 to 10 per cent of the total 
output of the plant. For an interesting discussion of evaporative con- 
densers the reader is referred to the admirable article by Oldham in the 
Proceedings of the Institute of Mechanical Engineers, 1899, and re- 
produced as a serial in Engineering (London), April 28 to June 30, 1899. 
The following test of a vertical cast-iron tube evaporative surface con- 
denser (Table 95) will give some idea of the performance of this type 
of condenser. This condenser consisted of two rows of 4-inch vertical 
cast-iron pipes connected at the top by U bends and at the bottom by 
cast-iron manifolds. A perforated iron trough distributes the water 
over the center of the bend and causes it to flow in a thin stream over 
the surface of the tubes. A wet-air pump is used for withdrawing the 
condensed steam and air. No fan is used for hastening evaporation. 

See Chapter XXV, for evaporative surface condenser calculations. 

Evaporative Condensers: Engr., Lond., May 5, 1889, pp. 432, 442, 447; Engineer- 
ing, May 19, 1899, p. 661, June 2, 1899, p. 721, June 30, 1899, p. 861; Trans. A.S.M.E., 
14-696; Power, Nov. 16, 1909; Prac. Engr. U. S., June, 1910, p. 346. 

TABLE 95. 

TEST OF A CAST-mON. VERTICAL-TUBE, EVAPORATIVE SURFACE CONDENSER 

NATURAL DRAFT. 



Date 

Weather 

Barometer 

Temperature of air 

Cooling surface, external 

Duration of trial, minutes 

Weight of steam condensed, pounds 

Boiler pressure 

Weight of water in circulation 

Weight of fresh water added 

Vacuum in condenser 

Initial temperature of circulating water 

Final temperature of circulating water 

Temperature of *' make up " water 

Temperature of water in hot well 

Weight of steam condensed per hour, pounds . . . 

Weight of water circulated per hour, pounds 

Weight of " make-up " water added per hour. . . 
Weight of steam condensed per square foot of 

cooling surface per hour 

Weight of "make-up" water per pound of steam 

Condensed, pounds 



Sept. 12 


Sept. 13 


Wet 


Fine 


29.8 


29.5 


? 


60 


272 


272 


99 


115 


800 


800 


60 


60 


1830 


1830 


600 


640 


23.36 


24.1 


117.5 


113.9 


128.4 


125 


58 


58 


136.5 


131.8 


485 


427 


6786 


? 


364 


334 


1.8 


1.54 


0.75 


0.80 



COl^DENSERS 



547 



247. Location and Arrangement of Condensers. — In the modern 
power house one sees two general arrangements of condensers and 
auxiliaries : 

1. The independent or subdivided system, in which each engine or 
turbine is provided with its own condenser, air and circulating pumps. 

2. The central system, in which the condensers and auxiliaries are 
grouped together. Ordinarily one condenser suffices for all engines. 

The Independent System. — The condenser is usually placed 
close to and below the engine so that all condensation may gravitate 



j;i^ :' T^ 



I pJ-e'c-tlbltix^V^ 



Dfsclrargei 



I 




Floor Lino 




Atmosphere 
Atmos|jlieric 
Relief 
Valve 



Fig. 321. Jot Condenser Located below Engine-room Floor, 

into it. Figs. 321 and 327 show an application of this system with jet 
condensers. Here each condenser receives its supply of cooHng water 
from a main injection pipe and discharges into a main overflow pipe. 
The exhaust pipe leading to the condenser is by-passed through a suit- 
able atmospheric relief valve to a main free exhaust header so that the 
engine may operate non-condensing in case the vacuum breaks or the 
condenser is cut out. The chief feature of this arrangement is its 
flexibility, as each unit is complete in itself and independent of the 
others. By far the greater number of central stations are equipped with 
independent condensers. 

Occasionally a jet condenser is located on the same level with the 



548 



STEAM POWER PLANT ENGINEERING 




' I ' I ' I ' I ' I ' I ' 1 ■ I Atmospheric 
Tfelief Valve 



Fig. 322. Jet Condenser Located above Engine-room Floor. 




^ 



b To Hot Well 



Exhaust 



Engine Receiver 



m 



Conc<^nser 



~^ 



2£ 




Discharge^ 



L^^Ufg g 



Fig. 32o. Surface Condenser Located below Engine-room Floor. 



CONDENSERS 



549 



engine or even above it, Fig. 322, but such a location should be avoided 
if possible, as it usually necessitates a larger number of bends and 
joints in the exhaust pipes than the basement arrangement, and in- 
creases the possibility of air leakage. If the exhaust pipe does not 
drain directly into the condenser, the lowest point in the piping should 
always be provided with a drip which should be opened when the engine 
is shut down, as condensation and leakage are apt to fill the pipe with 
water if the engine stands for any length of time. The end of the drip 
should be connected so that water cannot be drawn back through the 
drip pipe and into the engine 
cylinder. The length of exhaust 
pipe and particularly the num- 
ber of bends between engine and 
condenser should be kept as 
small as possible, otherwise the 
engine may not derive the full 
benefit of the vacuum in the con- 
denser. A case is recorded where 
the exhaust piping and appurten- 
ances in connection with a 5000- 
horsepower engine caused a drop 
of several inches in vacuum be- 
tween condenser and exhaust 
opening of the low-pressure cylin- 
der. (National Engineer, December, 1906, p. 10.) The wet-air pump 
must always be located below the condenser chamber so that the con- 
densation may gravitate to it. 

Fig. 323 shows the arrangement of a surface condenser with com- 
bined air and circulating pump in connection with a horizontal cross 
compound engine. The condenser and appurtenances are placed below 
the engine, thereby permitting the condenser to be closely connected 
to the engine. 

Fig. 324 shows the arrangement of a surface condenser in connection 
with a pumping engine. The condenser is placed in series with the 
pump suction. 

Several typical installations of surface condensers in connection with 
various forms of condenser auxiliaries are shown in Figs. 325 to 328. 

Central Systems. — In the central condensing systems the con- 
denser is located at any convenient point and the exhaust from all the 
engines piped to it. Any arrangement of condenser and auxiliary 
machinery may be adopted which will favor the lowest cost of installa- 
tion and expense of operation. Except where continuity of operation 




Fig. 324. Surface Condenser Installed 
in Connection with Pumping. 



550 



STEAM POWER PLANT ENGINEERING 




a 

Hi 
o 

t 

I 



d 

ft 

a 

o 
O 



CONDENSERS 



551 



Pomp Discharge 
to Condenser 



Discharge from CondensBe 




Water Supply fiom 
Cold Well 



Fig. 326. Surface Condenser with Leblauc Pumps. 



I 




Fig. 327. Wheeler Rectangular Jet Condenser with Centrifugal Tail Pump and 
Rotative Dry Vacuum Pump in Connection with a 10,000-kilowatt Steam Turbine. 



552 



STEAM POWER PLA.NT ENGINEERING 



is absolutely essential, only one circulating pump and one air pump are 
installed. This reduces the number of auxiliary pumps and appli- 
ances to a minimum, with a consequent decrease in first cost and main- 
tenance. With properly designed exhaust piping the condenser may 
be located at a considerable distance from the engine without undue 
loss of vacuum. 

Central condensers have found great favor in power plants in which 
the individual units are subjected to extreme variations in load, as in 



C.L. of 



'''' " .. I '\' — V 




FiG. 328. 



Longitudinal Elevation of the 50,000 sq. ft. Condenser for the Common- 
wealth Edison Co. 



rolling mills. At the works of the IlUnois Steel Company, South Chicago, 
111., one condenser takes care of the steam from 15,000 horsepower of 
engines in the rail mill, and another condenses the steam from the 
15,000 horsepower of engines in the Bessemer steel mill. A notable 
installation of this system in connection with street-railway work is 
in the power house of the Northwestern Elevated Company, Chicago, 
where a single condenser takes care of the exhaust steam of five engines, 
11,000 horsepower in all. Fig. 330 shows the general arrangement of 
this installation. 



CONDENSERS 



553 



For a comparison of the advantages and disadvantages of the inde- 
pendent and central systems see Engineering Magazine, October, 1900, 
p. 56; Engineering, London, June 23, 1899, p. 615; and Engineering, 
July 17, 1903. 

Condenser Auxiliaries. — The various types of condenser auxil- 
iaries and their power requirements are treated at length in paragraphs 
307 to 314. 




Fig. 329. Barometric Condenser with Centrifugal Water Injection and "Rotrex" 

Air Pump. 

248. Cost of Condensers. — The curves in Fig. 331 compiled by A. R. 
Smith of the Construction Engineering Department, General Electric 
Company, show the approximate costs of condensers including their 
auxiliaries, f.o.b. factory. The average for each capacity of turbine 
was compiled from costs without regard to surface, quality of water, 
vacuum maintained and steam or electric drive. Actual cost may vary 
considerably from those shown on the curves, depending on local con- 
ditions and other special considerations. 

The following figures give an idea of the relative costs of the different 
types of condensers and auxiUaries for a 1000-i.hp. plant using 20 pounds 
of steam per i.hp-hour at rated load, or a total of 20,000 pounds per 
hour. Vacuum to be maintained, 26 inches unless otherwise stated; 



554 



STEAM POWER PLANT ENGINEERING 




CONDENSERS 



555 



temperature of cooling water, 70 dcg. fahr.; hot-well temperature, 
105 to 120 deg. fahr.; distance between engine exhaust opening and 
mean level of intake well, 10 feet. 



S 10.00 



$9.00 



_^$8.00 
"3 

|$7.00 
O 

I $6.00 

S3 

H$5.00 

o 

^$4.00 
W 
u 
C.S3.00 

(C 

O 
'^§2.00 



$1.00 



































































































\ 
















































' 
















































\ 














































w 














































\ 


\ 














































\ 




s 














































1 


\ 














































V 




\ 












































\ 






X 


^ 








































• 


\ 








"^ 




^^ 


- 
































, 


\ 




' 










Sut 


TTc 




'Con.17 


L_ 




; 
















* 




V 


... 














































• 


<. 


"S 










































. 


• 


, 




^ 




.^ 




















































Jt 


t ( 


7on 


fj-^ 









__ 




































































































































_^ 
























L 



1000 2000 3000 4000 5000 6000 7000 8000 9000 10000 11000 12000 
Kw. Rating of Turbine 

Fig. 331. Curves Showing Approximate Cost of Condenser Equipment per Kilo- 
watt of Turbine Capacity. 
Siphon Condensers. 

1 16" siphon condenser with 6" centrifugal pump driven by 6" by 6" ver- 
tical engine $800 

Jet Condensers. 

1 14" by 22" by 24" jet condenser with single horizontal direct-acting pump 1335 

1 16" by 24" by 18" jet condenser with single vertical direct-acting pump 1620 

1 14" by 24" by 18" jet condenser with single vertical flywheel vacuum pump 1770 

1 12" by 17" by 22" by 25" jet condenser, single horizontal direct-acting 

compound pump 2200 

Barometric Condensers. 

1 barometric condenser, 10" by 16" by 12" horizontal single-cylinder 
rotative dry-air pump; 8" horizontal volute centrifugal pump 
direct connected to 23-horsepower high-speed engine 2500 

1 barometric condenser, 16" by 16" dry-air pump direct connected to 
9" by 16" steam engine; positive rotary pump, for circulating 

cooling water, belted to above engine 4300 

Surface Condensers. 

1 surface condenser, 1025 square feet cooling surface, mounted over 7^" 

by 14" by 14" by 12" combined air and circulating pump 2100 

1 surface condenser, 1025 square feet cooling surface, with 7|" by 12" by 
12" horizontal air pump, direct acting, and 6" centrifugal pump 
driven by 5" by 5" engine 2300 

1 surface condenser, 1025 square feet cooling surface; 5" by 12" by 10" 
Edwards single-cylinder air pump and 6" centrifugal pump driven 
by a 5" by 5" engine; maximum 28", referred to 30" barometer 2850 

1 surface condenser, 1025 square feet cooling surface; 6" by 8" rotative 
dry-air pump; 6" by 6" Edwards wet-air pump and 6" centrifugal 
pump driven by 5" by 5" engine; maximum vacuum 29", referred 
to 30" barometer (temp, cooling water 50 deg. fahr.) 3500 



556 STEAM POWER PLANT ENGINEERING 

Westinghouse-Leblanc Jet Condenser. 

1 jet condenser with turbine-driven pumps, 20,000 pounds steam per hour, 

26" vacuum, 70 deg. fahr. inlet water 2150 

1 jet condenser with turbine-driven pumps, 20,000 pounds steam per hour, 

29" vacuum, 50 deg. fahr. inlet water 3275 

In general the cost of complete condensing equipments installed 
and ready for operation will approximate as follows: 

Cost per Kilowatt of Main 
Generating Unit. 

Siphon condensers without air punp $2 . 00 to $3 . 00 

Jet condensers 3 . 00 to 4 . 50 

Barometric condensers with dry-air pump 4.00 to 6.00 

Surface condensers for 26-inch vacuum 3 . 50 to 5 . 00 

High-vacuum surface condensers 3 . 50 to 10 . 00 

Leblanc jet condensers and pumps 2.00 to 6.00 

349. Choice of Condensers. — The proper selection of condenser and 
auxiliaries for a proposed installation depends upon the conditions 
under which the plant is to be operated. These conditions vary so 
widely in practice that only a few of the more important factors will 
be considered. The principal advantages and disadvantages of the 
three types of water-cooled condensers are as follows:* 

Advantages Disadvantages. 

Surface Condenser. 

Re-use of condensate for boiler feed. First cost high. 

Re-use of condensate for ice production. Maintenance high. 

Readily adapted to the weighing of Requires considerable building space to 

condensate for tests. ' remove tubes. 

Slightly better vacuum obtainable. Acidulated water or water containing 

Advantage of low pumping head through foreign matter in large quantities 

siphon action. may preclude the use of surface con- 
Less chance of losing vacuum because densers. 

a drop in vacuum does not affect More head room necessary to obtain 

water supply. sufficient head on hot-well pump. 

Barometric Condenser. 

Condenser proper not costly, but piping Long exhaust pipe line to condenser 

to it is expensive. which entails high initial cost and 

No possibility of flooding turbine as in greater possibility of air leaks. 

the case of a low jet condenser. Loss of vacuum between turbine and 

Maintenance low. condenser, which may amount to ^ 

The use of acidulated water possible. inch or even more. 

Requires less circulating water than sur- As condenser cone generally extends 

face condenser. above roof, it does not lend itself to 

Requires little building space. economical station design when boiler 

Equipment simple. No hot-well pump room and turbine room are parallel and 

necessary and in some forms no vac- contiguous. 

uum pump is required. Waste of condensate. 

* A. R. Smith, General Electric Review. 



CONDENSERS 557 



Jet Condenser. 



Least expensive type of condenser. Failure of removal pump would flood 

Requires less building space. turbine. Protection is provided by a 

Equipment simpler because hot-well vacuum-breaking float valve. 

pump is not necessary. Waste of condensate. 

Requires less circulating water than 

surface condenser. 
Maintenance low. 
The use of acidulated water possible. 

Steam-driven condenser auxiliaries have been universally recom- 
mended in preference to motor drives because any disturbances on the 
electrical end will not affect the auxiliaries. For example, suppose a 
short circuit occurs on some outside feeder and the speed. and voltage 
is reduced sufficiently to let the condenser auxiliaries drop out. First, 
the loss of vacuum on the turbine will necessitate the immediate gener- 
ation of double the amount of steam, but the boiler room is not pre- 
pared for this emergency, and the only alternative is to reduce the load. 
Second, the vacuum pump has to be started, and, third, the circulating 
pump started and primed. The operations consume considerable time, 
especially with chaotic periods of interruption. Should there be two 
turbines on the line, the duration of interruption is doubled. 

The dry vacuum pump and hot well pump can conveniently be made 
motor driven because the motors are small and can be self-starting. 
An interruption of 30 minutes of the vacuum pump or one minute of 
the hot-well pump ought to show but httle effect on the vacuum. 

Motor-driven auxiliaries are very desirable, in that they are cheaper 
in first cost and maintenance; they obviate the use of considerable 
steam and exhaust piping and the expense of maintenance and radiation 
incident thereto; the motor speeds are conducive to high pump effi- 
ciencies, and they are easily started and require little attention when 
running. To enjoy these advantages without sacrificing continuity of 
service is possible by feeding the auxiliaries for each turbine off a sepa- 
rate auxiliary turbine driving an exciter and a-c. generator. This 
may seem like an additional complication, but investigation will show 
that this auxiliary turbine can be operated at a speed of highest economy, 
and each pump can be operated at the most eflftcient speed. The 
auxihary turbine can be exhausted into its own feed-water heater. 
See Fig. 357. 

Unless an auxiliary turbine is employed, or steam-driven auxiliaries 
used, there is usually a shortage of exhaust steam for heating the feed 
water. Take, for example, a case where the turbine is running at half 
rated load: the steam-driven exciter and boiler feed pump would be 
taking about 5 per cent of the total steam, which would heat the feed 



558 STEAM POWER PLANT ENGINEERING 

water from 75 deg. fahr. (29 in. vacuum) to 125 deg. fahr. If the main 
turbine were carrying full rated load, the condition would be worse, 
as the auxiliary steam would represent only about 3J per cent and the 
increase in feed water temperature would be only 35 deg. fahr. Now, 
if the condenser auxiharies are steam driven, the total exhaust steam 
would be about 15 per cent at half load and 9 per cent at full load, and 
the feed temperature in the former case would be 210 degrees with a waste 
of 1| per cent of the steam, and at full load it should be 165 deg. fahr. 

The Selection of Steam Turbine Condenser: A. R. Smith, The National Engr., June, 
1914, p. 351. 

350. Water-Cooling Systems. — When an ample supply of coohng 
water is unobtainable, for natural or economic reasons, the circulating 
water may be used over and over again by employing suitable cooling 
devices. The three most common in practice are 

1. The simple coohng pond or tank. 

2. The spray fountain. 

3. The cooling tower. 

251. Cooling Pond. — The water is cooled partly by radiation and 
conduction but principally by evaporation. The air is seldom satu- 
rated normally, and its capacity for absorbing moisture is increased on 
account of its temperature being raised by contact with the warm water 
and by radiation. The cooling action is independent of the depth of 
water and varies directly as the surface, the amount of heat dissipated 
for each square foot depending upon the temperature of the water, the 
relative humidity, and the velocity of the air currents. Results of 
tests are very discordant. 

Box in his Treatise on Heat states that the pond surface should ap- 
proximate 210 square feet per nominal horsepower for an engine work- 
ing twenty-four hours a day. (Treatise on Heat, Box, p. 152.) 

If the engine works only twelve hours per day, the area may be re- 
duced to 105 square feet per horsepower, because the water will cool 
during the night, but in that case the depth should be such as to give a 
capacity of 300 cubic feet per horsepower. These figures are based on 
a reduction in temperature of 122 to 82 deg. fahr., with air at 52 deg. 
fahr., and humidity 85 per cent, the steam consumption per nominal 
horsepower being taken at 62.5 pounds. It appears from tests that 
under ordinary conditions, in the northern part of the United States, 
with engines using 15 pounds of water per horsepower-hour and a 
vacuum of 26 inches, a reservoir having a surface of 120 square feet 
per horsepower would be ample for cooling and condensing water. 
(W. R. Ruggles, Proc. A.S.M.E., April, 1912, p. 607.) 



CONDENSERS 559 

Box gives the following formula for the rate of evaporation in per- 
fectly calm air: 

E = i243-\-S.7t){V -v), (230) 

in which 

E = evaporation in grains per square foot per hour, 
t = temperature of the water, deg. fahr., 

V = maximum vapor tension in inches of mercury at temp^ ^ure t, 
V = actual vapor tension. 

Evaporation is greatly affected by the force of the wind and varies 
from 2 to 12 times the amount determined from equation (230). 

Example 50. How many pounds of water will be evaporated per 
square foot per hour from a pond with the temperature of the water and 
air 80 deg. fahr.; air perfectly calm; barometric pressure 29.5 inches 
and relative humiaity 70 per cent? 

The maximum vapor tension at temperature of 80 degrees is 1.03 
inches of mercury. The actual vapor tension will be 

1.03 X 0.70 (= relative humidity) = 0.721. 

Substitute these values in equation (230). 

E = (243 + 3.7 X 80) (1.03 - 0.721) 
= 167 grains per square foot per hour 
= 0.024 pound per square foot per hour. 

A rough rule is to allow a heat transmission of 3.5 B.t.u. per hr. per 
sq. ft of pond surface per d/^gree fahr. difference in temperature between 
that of the air and water. 

253. Spray Fountain. — From equation (230) we see that even under 
the most favorable circumstances an enormous pond surface is neces- 
sary. To facilitate evaporation with a view toward reducing the size 
of the pond, the hot circulating water is sometimes distributed through 
pipes and discharged through nozzles, falling to the surface of the 
pond in a spray. 

The water issuing from the nozzles creates a draft which aided by 
the natural breeze, effects the necessary evaporation. The loss of 
water due to evaporation seldom exceeds 4 per cent of the weight of 
water circulated. The pressure required at the nozzles is approxi- 
mately 6 pounds per sq. in. and in many cases the condenser pump is 
able to furnish the necessary pressure. Under ordinary conditions the 
power necessary to operate the sprays will average less than \\ per cent 
of the power generated by the prime mover. Should the temperature 
of the condenser discharge water exceed the limit of reduction by single 
spraying the desired reduction in temperature may be effected by double 
spraying. In this arrangement the condenser discharge is mixed in 



560 



STEAM POWER PLANT ENGINEERING 



the hot well with an equal amouni: of cooler water flowing through an 
equalizing valve from the spray pond. The resulting mixture is pumped 
to the nozzles and resprayed. Some idea of the performance of a 
spray cooling system may be gained from the data in Tables 95 and 96. 



TABLE 96. 

SINGLE-SPRAY SYSTEM — 6000-KW. STEAM TURBINE PLANT. 







Temperatures, Decrees Fahrenheit. 




Month. 


Humiditv, 
Per Cent. 












S A.M. 


12 m. 


4 P.M. 


t 

Remarks. 


Jan 


62 


( Discharge water. . . 

< After spraying 

( Surrounding air... . 


68 

48 
8 


73 
53 
14 


73) 
53 > 
20) 


Clear 


Mar 


50 


( Discharge water . . . 

< After spraying 

( Surrounding air. . . 


79 

58 
30 


86 
66 
50 


90) 
70 [ 
43) 


Clear 


May 


72 


C Discharge water. . . 

< After spraying 

( Surrounding air 


89 
70 
65 


94 
75 

72 


97) 
78 > 
70 ) 


Clear 


July 


70 


C Discharge water . . . 

< After spraying 

( Surrounding air. ... 


108 
90 
90 


118 
93 

98 


118 ) 

93 > 

102) 


Clear 


Aug 


84 


C Discharge water . . . 

< After spraying 

( Surrounding air 


112 

88 
72 


114 

89 

74 


116) 
90 > 
79) 


Cloudy 


Nov 


70 


C Discharge water . . . 

< After spraying 

( Surrounding air 


89 
62 

27 


90 
64 
33 


88) 
63 J 
34) 


Cloudy 



TABLE 97. 
DOUBLE-SPRAY SYSTEM. 



First Spraying. 



Second Spraying. 



Temperature air, deg. fahr 

Relative humidity, per cent 

Temperature hot water, deg. fahr 

Temperature, cooled water, deg. fahr. 
Total degrees cooled, fahr 



87.0 

48.5 

122.5 

88.3 



88.0 
46.0 
88.7 
78.8 
44.1 



Natural ponds without sprays require about 50 times more area 
than spray cooling systems. A rough rule is to allow 130 B.t.u. per 
sq. ft. per hr. per degree difference in temperature. 

253. Cooling Towers. — A cooling tower consists of a wooden or sheet 
iron housing open at the top and bottom and so arranged that the 



CONDENSERS 



561 



OISTRIBUTINO 
TROl/CH 



hot water may be elevated to the top and distributed in such a manner 
that it falls in thin sheets or sprays into a reservoir at the bottom, air 
at the same time being drawn in at the bottom by natural draft or 
forced in by a fan. The water gives up its heat to the ascending cur- 
rent of air by evaporation, convection and radiation, the latter, however, 
being a relatively small fac- 
tor. Of these, evaporation 
absorbs from 75 to 85 per 
cent of the heat, convection 
or direct transfer of heat to 
the air comes next, while 
radiation partly in the tower 
and partly through the pip- 
ing accounts for the balance. 
If the air supply is dependent 
entirely upon the chimney 
action of the device the sys- 
tem is known as a natural 
draft or flue cooling tower; 
if the air is forced into the 
device by fans the system is 
called a forced draft cooling 
tower. Water cooling towers 
may be classified as (1) forced 
draft, (2) natural draft — 
open type or atmospheric, 
(3) natural draft — closed or 
flue type, and (4) combined 
forced and natural draft. 

Forced draft towers are 
completely enclosed, except 
at the top and at the base 
where provision is made for 
the fan openings. In the at- 
mospheric type of natural 
draft tower the sides are 
louvered and the necessary ^^^- ^^2. Barnard-Wheeler Cooling Tower. 

air is supplied through the open base and through the louvered sides 
by natural air currents. The flue type of natural draft tower receives 
its air supply through the chimney action of the flue. The combined 
forced and natural draft tower may be used with natural draft only 
for light loads and forced draft for heavy loads. 




DISCHARGE 
FROM 

TOWCR 



562 



STEAM POWER PLANT ENGINEERING 



The different designs vary principally in the method of water dis- 
tribution. Fig. 332 illustrates the Barnard- Wheeler cooHng tower 
in which the falling water is broken up by vertically suspended gal- 
vanized iron wire cloth mats, causing it to trickle in thin sheets to the 
bottom. A similar result is brought about in the Worthington tower 
by pieces of terra cotta pipe 6 inches in diameter and two feet long 




Fig. 333. C. H. Wheeler Atmospheric Cooling Tower. 

placed on ends in rows. In the standard type of Alberger cooling tower 
the water trickles down the sides of swamp-cypress boards arranged in 
honeycomb fashion. In the Alberger improved type the fan is placed 
at the top of the tower with its shaft in a vertical position. The fan is 
operated by a Pelton water wheel which receives its power from a tur- 
bine pump. No oil lubrication is employed, and the operating mecha- 
nism is controlled entirely from the engine room. In the Jennison 
cooling tower the water is divided into a rain of drops, constantly re- 



CONDENSERS 



563 



tarded in their fall by a scries of perforated 4 X 4-inch galvanized-iron 
trays arranged in horizontal rows and staggered vertically. 

With the best forms of cooling towers, under average conditions, the 
temperature of the circulating water may readily be reduced from 40 
to 50 degrees with a loss not exceeding 3 or 4 per cent of the total quan- 
tity of water passing through the tower. The power consumed by the 
fan in a forced-draft apparatus averages 2 per cent of that developed 
by the main engines, for the maximum requirements during summer 
months, and 1} per cent during the winter. 

The location of the tower may be on the engine-rooom floor, on top of 
the building, or in the yard, the latter being the most adaptable. It 
may be any reasonable distancie from the engine and condenser. 

354. Test of Cooling Towers. 

RESULTS OF TEST OF NATURAL-DRAFT TOWER, DETROIT. 

Complete Five-Fifths Surface Installed. 

Proc. A.S.M.E. Mid- Nov., 1909, p. 1205. 

Engines: Two 400-i.hp., 300-kw. Macintosh & Seymour tandem-compound 

engines, overhung generators. 

Condensers: Worthington surface (admiralty type) IQOO-sq. ft. reciprocating wet- 
air pump and circulating pump. 

Tower: Wood-mat construction, 24,500 sq. ft. evaporating surface, exclusive 

of shell. 

Test: March 15 to 16, 1901, 4 p.m. to 4 p.m., 24 hr. 

A.M. P.M. Average. 

Weather: Barometer (abs.), min 30.22 30.07; 30.14 30.27 

Temperature air, deg 18.5 25; 30 25 

Relative humidity, per cent. . . 76 82; 58 72 

Load: 600 kw. max. to 50 kw. min. Average 244.9 kw. 

Engine efficiency = 92.5 = 875 i.hp. max. Average. . . . 354.8 i.hp. 

Steam: Weight of condensed steam per hr., lb 5910. 6 

Temperature exhaust steam, deg. fahr 134. 38 

Temperature condensed steam, deg. fahr 108. 78 

Weight of steam per hour, max. load, lb 13,500 

Vacuum (abs.) 25 to 19, average about 22 

Vacuum corresponding to temperature exhaust steam .... 25 

Vacuum possible with good condenser (10 deg. difference) 28 

Water: Circulated per hr., lb 293,536 

Temperature hot well, average, deg. fahr 87 . 50 

Temperature cold well, average, deg. fahr 71 . 27 

Vaporization loss per hr., lb 5970 

Results: Condenser surface per kw., sq. ft 2. 66 

Steam per kw-hr., lb 24 . 3 

Steam per i.hp-hr., lb 16. 66 

Circulating water per lb. of steam, lb 49. 6 

Steam per sq. ft. condenser surface per hr., lb 3.7 

Circulating water per sq. ft. tower surface, lb 12 

Difference in temperature between exhaust steam and dis- 
charge, deg. fahr 47 



564 



STEAM POWER PLANT ENGINEERING 



Cooling: Max. 20 deg., min. 3 deg.-5 deg. Average 16.23 

Heat dissipated per hr., B.t.u 4,769,000 

Heat per sq. ft. tower surface, B.t.u 195 

Heat per sq. ft. per 1000 lb. water, B.t.u 0. 665 

Evaporation : Circulating water, per cent 2 . 03 

Engine steam, per cent 101 

Tower: Surface per kw. (average load 245 kw.), sq. ft 100 

Surface per kw. (max. load 600 kw.), sq. ft 40. 8 

Surface per 1000 lb. steam max. load, sq. ft 1820 

Surface per 1000 lb. steam average load, sq. ft 4140 

Surface per 1000 lb. circulating water per deg. max. cooling, 

sq.ft 4.17 





Temperature, Deg. Fahr. 


Quantities. 


Time. 


Air. 


Hot 
Well.* 


Cold 
Well. 


Water 
Cool- 
ing. 


Total 

Heat 

Head.t 


Tower 

Water, Lb. 

per Hr. 


Heat Dissi- 
pated, B.t.u. 
Lb. per Hr. 






Load, 
Kw. 


1 


2 


3 


4 


5 


6 


7 


8 


9 


10 


11 


12 noon 

1.30 

2.30 
3.30 

4.30 
5.00 
6.00 
7.00 
8.00 


34 

35 

35 
35 

32.5 

28.5 

26 

24 

24 


102 

106.5 

106.5 
113 

100 

103.5 

125 

121 

123 


89 

90 

87.5 
88.5 

84 
88 
94 
94 
94.5 


13 

16.5 

19 
24.5 

16 

15.5 

31 

27 
28.5 


68 

71.5 

71.5 
78 

67.5 

75 

99 

97 

99 


375,000 

1(375,000 

* 1 370,200 

375,000 

375,000 

399,000 
445,500 
417,000 
427,000 
427,000 


4,880,000 

6,108,000 

7,120,000 
9,000,000 

6,384,000 

6,900,000 

12,930,000 

11,532,000 

12,174,000 


332 

415 

484 
613 

434 
470 

880 

785 
827 


25 

24.8 

25 
25 

26.6 

29.7 
27.8 
27.4 
27.4 


270 

(315 

1290 

315 

350 

365 

485 
655 
570 
600 



* Assuming a more efficient condenser, say 10 deg. difference, the probable vacuum would be 
26 deg. to 27.5 deg. This condenser actually operated at 40 deg. to 50 deg. difference. 

t Total heat head = air heatmg + lost head. X Difference due to rapid change in load. 

For cooling tower calculations and problems in hygrometry see Chapter XXV. 



PROBLEMS. 

1. Reading of vacuum gauge 26.5, temperature of room 80 deg. fahr., barometer 
29.5, temperature of mercury in the barometer 40 deg. fahr. Determine the vacuum 
referred to a 30-inch barometer. 

2. If the absolute temperature in a condenser is 5 inches of mercury and the 
temperature of the air-vapor mixture in the chamber is 90 deg. fahr., required the 
percentage of air (by weight) in the mixture. 

3. If the temperature within a condenser is 100 deg. fahr. and there is entrained 
0.1 lb. of air per lb. of steam, required the maximum degree of vacuum obtainable. 

4. Required the volume of aqueous vapor to be withdrawn in order to cool 10,000 
lb. of water from 120 to 80 deg. fahr. 

5. A 30,000-kw. turbine uses 12 lb. steam per kw-hr., initial pressure 290 lb. abs., 
superheat 250 deg. fahr., vacuum 28.5 in. referred to a 30-in. barometer; initial 
temperature of the cooUng water 70 deg. fahr., water velocity through tubes 8 ft. 
per sec. Required: 

a. Weight of cooling water. 

b. Sq. ft. condenser tube surface. 



I 



CONDENSERS 565 

c. Number of 18 B.W.G. tubes in each pass of the condenser. 

d. Length of water travel. 

6. A 200-kw. turbine uses 20 lb. steam per kw-hr., initial pressure 150 lb. absolute, 
superheat 100 deg. fahr., vacuum 27 in. referred to a 30-in. barometer. If an evap- 
orative surface condenser of the forced draft type is used to create the vacuum, 
required the amount of atmospheric air and water spray which must be forced 
through the condenser. The temperature of the atmospheric air is 80 deg. fahr., 
wet bulb thermometer 65 deg. fahr., air issuing from the condenser is completely 
saturated and its temperature is 15 degrees below that of the vapor in the condenser, 
fan pressure 4 in. of water. 

7. How much "make up" water is necessary for the cooling tower system of a 
steam engine plant operating under the following conditions: Engines 1000 hp., 
water rate 20 lb. per i.hp-hr. initial pressure 120 lb. abs., vacuum 26 in., barometer 
30 in.; temperature of injection water, discharge water and atmospheric air, 90, 
110 and 70 deg. fahr., respectively; relative humidity of air entering and leaving 
tower 65 and 95 per cent respectively. 



CHAPTER XII 

FEED WATER PURIFIERS AND HEATERS 

255. General. — All natural waters contain more or less foreign 
matter either in suspension or solution. The organic constituents of 
this foreign matter are of vegetable and animal origin taken up by 
vvater flowing over the ground or by direct contamination with sewage 
and industrial refuse. Feed water containing organic matter may 
cause foaming due to the fact that the suspended particles collect on 
the surface of the water in the boiler and impede the liberation of the 
steam bubbles arising to the surface. 

The suspended inorganic impurities consist of clay, silica, iron, alu- 
mina, and the like, in the form of mud and silt. The more common 
soluble inorganic impurities are lime, magnesia, iron and sodium in the 
form of carbonates, sulphates and chlorides, oxides of silica, iron and 
alumina, some free carbonic acid and occasionally free sulphuric acid 
and hydrogen sulphide. 

When raw water is fed into a boiler all of the solids remain in the 
boiler and are constantly increased in amount by the evaporation 
taking place. Some of the accumulated impurities deposit on the 
heating surface as scale, some are present as suspended matter and 
others remain in solution. The most widely known evidence of the 
presence of scale-forming ingredients in feed water is known as hardness. 
If the water contains only such ingredients as carbonates of lime, 
magnesia and iron which may be precipitated by boiling at 212 deg. 
fahr. it is said to have temporary hardness. Permanent hardness is due 
to the presence of sulphates, chlorides and nitrates of lime, magnesia 
and iron which are not completely precipitated at a temperature of 
212 deg. fahr. Hardness is conveniently determined by means of a 
standard soap solution as follows: 

A 100-cc. (cubic centimeter) sample of water to be tested is put in a 
250-cc. bottle and a standard soap solution (this may be obtained from 
chemical dealers) run in 0.2 cc. at a time, the bottle being shaken vigor- 
ously after each addition of the soap solution. Finally a lather is 
produced that will persist for at least five minutes, and then the volume 
of soap solution used in cc. gives the degrees ''U. S." hardness. One 
degree '^U. S." hardness is equivalent to 1 grain of calcium carbonate 
per U. S. gallon (1 part in 58,349). 

566 



FEED WATER PURIFIERS AND HEATERS 



567 



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568 STEAM POWER PLANT EXGINEEPJNG 

The following factors may be used for specifying hardness of water 
in terms of calcium carbonate per U.S. gallon. 
Magnesium carbonate X 1.19 



Magnesium sulphate X 0.833 

Calcium sulphate X 0.735 

Magnesium chloride X 1.05 

Calcium chloride X 0.901 



= hardness as calcium carbonate, 
grains per U. S. gallon or U. S. 
degrees. 



It is impossible to judge the quality of feed water merely by the 
grains of solids per gallon since a large amount of soluble salt such as 
sodium chloride will not be as deleterious as a very small amount of 
calcium sulphate. 

The scale of hardness usually accepted (grains of dissolved salts per 
U. S. gallon) is as follows: Soft water, 1 to 10; moderately hard 10 to 
to 20; very hard water, above 25. 

The following is a rough rating according to the number of grains of 
incrusting sohds per United States gallon : 

Less than 

8 grains very good. 

12 to 15 grains good, 

15 to 20 grains fair. 

20 to 30 grains bad. 

Over 30 grains very bad. 

This apphes to calcium carbonate, magnesium carbonate, and mag- 
nesium chloride. For water containing sulphate of calcium and mag- 
nesium, divide the first column by 4 for the same rating. 

The limiting factor in deciding whether a water carrying a large 
amount of soluble salts may be used for boiler feed purposes is the 
amount of blowing down necessary to keep the degree of concentration 
within the limits found by experience. 

256. Scale. — Scale is formed on boiler heating surfaces by the depos- 
iting of impurities in the feed water and varies from a porous, friable 
crust to a dense, very hard coating. The amount of scale formed does 
not bear a direct relation to the amount of impurities present but de- 
pends on the type of boiler, rate of driving and the nature of the scale- 
forming ingredients in the water. Scale tends to lower the efficiency 
and capacity of the boiler and may cause overheating of the plates 
and tubes. 

Table 99 gives the results of a number of tests made on locomotive 
boiler tubes with different thicknesses and characters of scale. The 
diversity of the results indicates the futility of basing the decrease in 
conductivity on the thickness of the scale. For example, test No. 1 
shows a decrease in conductivity of 9.1 per cent for a scale 0.02 inch 
thick, while No. 16 shows a decrease of only 6.75 per cent for a scale 



FEED WATER PURIFIERS AND HEATERS 



569 



over 6.5 times as thick. The scale in each case was even, hard, and 
dense. Again, No. 8 with a very soft scale 0.042 inch thick gives a 
decrease in conductivity of 9.54 per cent, whereas No. 14, also very 
soft but twice as thick, gives a decrease of only 4.95 per cent. No doubt 
the heat transmission is a function of the chemical as well as the physical 
properties, but further experiments are necessary before any specific 
conclusion can be drawn. 

TABLE 99. 
INFUENCE OF SCALE ON HEAT TRANSMISSION. 
(Locorriotive Boiler Tubes.) 



No. 


Thickness of Scale, 
Inches. 


Character of Scale. 


Decrease in Con- 
ductivity due to 
Scale. Per cent. 


1 


.02 

.02 

.033 

.033 

.038 

.04 

.04 

.042 

.047 

.065 

.07 

.07 

.085 

.089 

.11 

.13 


Hard, dense 

Hard 

Soft 

Very hard 

Medium 

Soft, porous 

Hard, dense 

Very soft 

Hard 

Medium 

Soft 

Hard 

Soft, porous 

Very soft 

Hard, porous 

Hard, dense 


9.1 


2 


2.02 


3 


4.3 


4 


3.5 


5 


4 03 


6 


6 82 


7 


3 07 


8 


9 54 


9 

10 


2.75 
2 39 


11 


2.38 


12 


4 43 


13 


19 


14 


4 95 


15 


16 73 


16 


6 75 







From tests conducted at the University of Illinois, Railroad Gazette, Jan. 27, 1899, June 14, 1901. See 
also Engineering Record, Jan. 14, 1905, p. 53; Power, February, 1903, p. 70; Street Railway Review, July 15, 
1901, p. 415. 

A moderate amount of scale has little influence on the efficiency and 
capacity of boilers operating at or below normal rating but for high 
driving rates scale must not be permitted to accumulate. In the modern 
central station with its heavy peak loads pure feed water is of vital 
importance to economy and continuity of operation. For scale preven- 
tion see paragraphs 260 to 266. 

257. Foaming and Priming. — Boiler troubles due to foaming or prim- 
ing are often caused by concentration of alkali salts in the water 
within the boiler, although silt, organic matter, loosened scale, lubri- 
cating oil, rate of driving and the design of the boiler all have bearing 
upon this phenomenon. Where this is caused by excessive concen- 
tration it may be largely overcome by frequent blowing down. Sur- 
face blowing is, of course, a remedy where it can be applied. Foaming 
caused by organic matter in suspension may be minimized by filtration. 



k 



570 



STEAM POWER PLANT ENGINEERING 



258. Internal Corrosion. — Corrosion is evidenced by small pits or 
depressions and by large cup-shaped hollows on the metal surface, and 
occasionally by a considerable destruction of a large portion of the 
surface. Carbonic acid gas, occluded oxygen, sodium, calcium and 
magnesium chlorides are common causes of corrosion. Magnesium and 
calcium chlorides are very pernicious in that they produce free hydro- 
chloric acid on hydrolysis. Corrosion is also found in boilers using a 
high percentage of condensate or distilled water. A theory* accepted 
by physicists embraces the fact that in the presence of a solvent the 
iron goes into solution as a hydrate before oxidizing. Considering 
that water is a universal solvent every metal has an inherent tendency 
to dissolve in water or water solutions. This tendency is called the 
solution tension of the metal. Opposing this tendency to dissolve is a 
pressure in the solution tending to resist the entrance into the solution 
of any more of the metal. This opposing pressure is known as the os- 
motic pressure of the solution. Accepting this theory, it is only neces- 
sary, in order to prevent corrosion, to raise the osmotic pressure of the 
solution or electrolyte, above the solution tension of the metal. In 



TABLE 100. 

SUMMARY OF INSPECTOR'S REPORTS FOR THE YEAR 1916. 
(Hartford Steam Boiler Inspection and Insurance Company.) 



Nature of Defects 



Whole Number. 


Dangerous. 


28,212 


1,593 


42,877 


1,612 


2,568 


315 


19,008 


793 


10,968 


814 


984 


266 


2,049 


504 


9,401 


814 


3,711 


563 


5,361 


498 


278 


27 


1.448 


201 


12,554 


1,581 


15,080 


4,989 


5,537 


373 


4.192 


714 


4,262 


1,337 


420 


123 


1,386 


235 


1,695 


358 


8,351 


815 


32 


32 


4,261 


662 


184,635 


19,219 


25.901 





Cases of sediment or loose scale 

Cases of adhering scale 

Cases of grooving 

Cases of internal corrosion 

Cases of external corrosion 

Cases of defective bracing 

Cases of defective staybolting. . 

Settings defective 

Fractured plates and heads 

Burned plates 

Laminated plates 

Cases of defective riveting 

Cases of leakage around tubes. . 
Cases of defective tubes or flues 

Cases of leakage at seams 

Water gages defective 

Blows-offs defective 

Cases of low water 

Safety valves overloaded 

Safety valves defective 

Pressure gages defective 

Boilers without pressure gages. . 

Miscellaneous defects 

Total 

Condemned 



* A. H. Babcock, Trans. A.S.M.E., Vol. 37, p. 1119. 



FEED WATER PURIFIERS AND HEATERS 



571 



the case of boilers, the osmotic pressure of the electrolyte or boiler 
water, is raised by the addition of alkahne salts. The osmotic pressure 
being dependent on the concentration of the salts, the corrosive con- 
dition of the electrolyte is indicated by its alkaline strength. The 
alkaline strength to be carried in the boiler depends on the salts used. 

Table 100, compiled by the Hartford Steam Boiler Inspection and 
Insurance Company, shows the number of boilers inspected by that 
company during the year 1916 and the number found defective from 
various causes. 

259. General Feed Water Treatment. — Table 101 (''Boiler Waters," 
W. W. Christie) outlines some of the troubles arising from feed water, 
their cause and means for preventing them. 

TABLE 101. 

BOILER TROUBLES ARISING FROM USE OF IMPURE FEED WATER. 



Trouble. 



Incrustation. . 



Priming. 



Corrosion. . . . - 



Cause. 



Sediment, mud, clay, etc.. . \ 
Readily soluble salts 

Bicarbonate of magnesia, J 
lime, iron j 

Organic matter 

Sulphate of lime < 

Organic matter 

Grease < 

Chloride or sulphate of ) 

magnesium ) 

Sugar } 

Acid ( 

Dissolved carbonic acid and 
oxygen 

Electrolytic action 

Sewage ■< 

Alkalies 

Carbonate of soda in large ) 
quantities ( 



Remedy or Palliation. 



Filtration. 

Blowing off. 

Blowing off. 

Heating feed and precipitate. 

Caustic soda. 

Lime. 

Magnesia. 

See below. 

Sodium carbonate. 

Barium chloride. 

Precipitate with alum ) 

Precipitate withferric > 



recipitate' 
chloride 
Slaked lime 
Carbonate of sod 

Carbonate of soda. 



a[ 



and filter 



and filter 



Alkali. 

Slaked lime. 

Caustic soda. 

Heating. 

Zinc platen. 

Precipitation with alum or ferric 

chloride and filter. 
Heating feed and precipitate. 

Barium chloride. 



The neutrahzation or elimination of the impurities may be effected 
by one or more of the following methods: 
1. Chemically. 

Water-softening plants. 
Boiler compounds. 



572 STEAM POWER PLANT ENGINEERING 

2. Mechanically. 

Filters. 
Blow-off. 
Tube cleaners. 

3. Thermally. 

Feed-water heater. 
Distillation. 

360. BoUer Compounds. — The object of treatment with boiler com- 
pounds is to neutralize the evil effects of the impurities in the feed water 
or to change them into others which are less objectionable and which 
are easily removed. When properly compounded and introduced into 
the boiler such preparations are of great benefit, but when improperly 
used they may produce even greater troubles than the impurities 
which they are expected to eliminate. 

Boiler compounds may be divided into three classes: 

1. Those converting the scale-forming elements into new substances 
which will not form a hard, resisting scale and which are readily removed 
by skimming, blowing off, or by tube cleaners. For example, feed 
water containing sulphates of lime and magnesia will form a dense, 
tenacious scale. If carbonate of soda be added in correct amount the 
sulphates are converted into insoluble carbonates which are precipi- 
tated and form scale varying from a more or less porous, friable crust to 
a soft "mush" or mud. The resulting sulphate of soda remains in 
solution and does not form scale unless allowed to concentrate and this 
is prevented by blowing off. An excess of soda is apt to cause foaming 
and at high temperatures is liable to attack the inside of gauge glasses. 
Bisodium and trisodium phosphate, sodium tannate, fluoride of so- 
dium, sugar, etc., have all proved satisfactory, but as each case requires 
special treatment no detailed discussion is possible within the scope 
of this work and the reader is referred to the accompanying bibhography. 

2. Those enveloping the newly precipitated scale-forming crystals 
with a surface which prevents them from cementing together. The 
ingredients used to bring about this result are starches, woody fibers, 
dextrine, slippery elm, and the like. 

3. Those preventing the formation of hard scale by a solvent or 
"rotting" action, as kerosene and petroleum oils. 

Under favorable conditions all that the most effective boiler compound 
can do is to change the nature of the precipitate from one which adheres 
to the boiler to one which will be carried in suspension. The accumu- 
lation of sludge in the boiler resulting from the use of a compound can- 
not be entirely removed by blowing off and consequently frequent 
washing out becomes a necessity. 



FEED WATER PURIFIERS AND HEATERS 573 

Compounds for miniinizing the formation of scale are recommended 
for use only in small plants where the cost of treating the water before 
it enters the boiler is prohibitive or in plants where space limitations 
prevent the installation of a purifying plant. 

Patented Boiler Compounds: Prac. Engr., Aug., 1911, p. 523. 

261. Use of Kerosene and Petroleum Oils in Boiler Feed Water. — 

Kerosene oil and other refined petroleum oils are sometimes used with 
good effect in boilers to soften scale. These oils are said to change the 
deposit of lime from a hard scale to a friable material which may be 
easily removed. To be reasonably effective the kerosene should be in- 
troduced after the boiler is emptied and washed and the refilling should 
be effected from the bottom. Kerosene should not be fed into the 
boiler with the feed water since it may form a non-conducting film over 
the heating surfaces. 

Use of Kerosene in Boilers: Engr. U. S., Sept. 15, 1905, p. 634; Eng. News, May 
24, 1890, P.M97; Powsr, Nov. 8, 1910, p. 1993; Trans. A.S.M.E., 9-247, 11-937; 
Locomotive, July, 1890, p. 97. 

262. Use of Zinc in Boilers. — Zinc is often introduced into boilers 
to prevent corrosion. The theory is that a feeble but continuous cur- 
rent of hydrogen is generated over the whole extent of the iron by 
electrolytic action. The bubbles of hydrogen formed isolate the 
metallic surface from scale-forming substances. If there is but a little 
of the scale-forming element it is precipitated and reduced to mud; if 
there is considerable, coherent scale is produced which takes the form of 
the iron surface but does not adhere to it, being prevented from doing 
so by the intervening bubbles of hydrogen. Zinc is ordinarily sus- 
pended in the water space of the boiler in the shape of blocks, slabs, 
or as. shavings in a perforated vessel. Electrical connection between 
the metallic surfaces is essential. Rolled zinc slabs 12 X 6 X ^ inches 
have found much favor in marine practice. Generally speaking one 
square inch of zinc surface is sufficient for every 50 pounds of water 
in the boiler, though the quantity placed in the boiler should vary with 
the hardness. The British Admiralty recommends the renewing of the 
zinc slabs whenever the decay has penetrated to a depth of J inch 
below the surface. Zinc does not prevent corrosion or scale formation 
in all cases and may even aggravate the trouble. 

Use of Zinc in Boilers: Prac. Engr., Dec, 1911, p. 835; Power, Oct. 18, 1910, 
p. 1874; Sept. 27, 1910, p. 1734. 

263. Methods of Introducing Compounds. — Boiler compounds may 
be introduced into the boiler continuously or intormittentl^^ Small 
quantities introduced continuously or at short intervals are more effec- 



574 STEAM POWER PLANT ENGINEERING 

tive than large quantities at long intervals. Continuous feeding is 
ordinarily brought about by connecting the suction side of the feed 
pump with a reservoir containing the compound in solution, arranged 
similarly to an ordinary cyhnder oil lubricator. In large plants an 
independent pump is often used to force the solution into the feed line. 
Intermittent feeding is brought about by temporarily connecting the 
suction of the feed pump with the reservoir containing the compound. 
The use of boiler compounds does not necessarily prevent scale from 
forming in time, though it will reduce the evil to a minimum. In some 
instances where compounds are used it is found necessary to run a 
tube cleaner through the tubes at certain intervals, in others such a 
course has not been found necessary. 

364. Mechanical Purification. — Waters containing sand, mud, or- 
ganic matter, and in fact all matter which is not in solution or in chemi- 
cal combination with the water may be purified by mechanical filtra- 
tion. Mud and sand may be eliminated by simply permitting the 
water to stand for some time in setthng tanks. Suspended matter 
which will not gravitate to the bottom may be removed by filtering the 
water through coke, cloth, excelsior, or the hke. Filters should be in 
duplicate for continuity of operation. 

Vegetable and other organic impurities commonly float on the sur- 
face of the water when the boiler is making steam, and may be blown out 
through a "surface blow-out." (See paragraph 88.) 

Precipitated matter may be ejected from the boiler by frequent 
blowing off before it has time to adhere and bake to a crust. This 
procedure is particularly essential when boiler compounds are used. 

For description and use of mechanically operated tube cleaner see 
paragraph 92. 

265. Tliermal Purification. — (See also Live Steam Purifiers, para- 
graph 298.) The carbonates of lime and magnesia are held in solution 
in fresh water by an excess of carbon dioxide and are completely pre- 
cipitated by boiling. At ordinary temperatures carbonate of lime is 
soluble in approximately 20,000 times its volume of water, at 212 
deg. fahr. it is slightly soluble, and at 290 degrees it is insoluble. Sul- 
phate of lime is much more soluble in cold than in hot water, and is 
completely precipitated at 290 degrees. (Revue de Mecanique, Novem- 
ber, 1901, pp. 508, 743.) 

Thus it will be seen that the application of heat will completely pre- 
cipitate these scale-forming elements provided the temperature is high 
enough and sufficient time is allowed for action. In the commercial type 
of exhaust and live steam heaters complete precipitation cannot be 
effected on account of the short time the water is held in them, and be- 



i 



FEED WATER PURIFIERS AND HEATERS 575 

cause of the limited space for retaining the scale. There is no question 
but that some of the scale-forming elements are removed from the feed 
water by exhaust and live steam heaters but the amount precipitated is 
but a small fraction of the total except in cases of unusually pure water. 
Efficiency of Live Steam Feed Heater: Power, Feb. 21, 1911, p. 295. 

366. Water Softening. — When feed water contains a large amount 
of scale-forming material it is usually advisable to '^ soften" it before 
allowing it to enter the boiler rather than to introduce the chemical 
reagent into the boiler. The complete softening of water requires the 
removal of both its temporary and its permanent hardness. When 
water is softened outside the boiler and the sludge removed by sedimen- 
tation and filtration before deUvering it to the heater the chemicals 
used are almost invariably hme, Ca(0H)2, and soda ash, Na2C03, alone 
or in combination with each other. Other chemicals may effect the 
desired result more efficiently but their cost is prohibitive. The chemi- 
cal changes which take place when these reagents are added to water 
containing calcium sulphate, CaS04, magnesium sulphate, MgS04, 
calcium bicarbonate, Ca(HC03)2 or magnesium bicarbonate, Mg(HC03)2, 
are as follows: 

CaS04 + NaoCOa = CaCOa + Na2S04. (230) 

MgS04 + Ca(0H)2 + NasCOa = Mg(0H)2 + CaCOa + Na2S04. (231) 

Ca(HC03)2 + Ca(0H)2 = 2CaC0a + 2H2O. (232) 

Mg(HC03)2 + 2Ca(OH)2 = Mg(0H)2 + 2CaC0a + 2H2O. (233) 

From these reactions the amount of reagent to be added to raw 
water may be calculated by considering the combining weights as 
follows : 

For soda ash and calcium sulphate 

CaS04 : Na^COa = I \ x, (234) 

40 + 32+4 (16) : 2(23) + 12 + 3(16) = I \ x, (235) 

X = 0.779, 
in which 

X = soda-ash factor or the weight of soda ash required per lb. of cal- 
cium sulphate. 

By similar calculations the factors for salts which require soda ash 
are found to be as follows: 



Salt. 



Calcium chloride, CaCl2 0.955 

Magnesium chloride. MgCl. 1.113 

Magnesium sulphate, MgS04 ' 0.881 

Calcium sulphate, CaS04 . 779 



Soda-ash factor. 



576 STEAM POWER PLANT ENGINEERING 

For salts which require Hme : 





Factor. 


Salt. 


Lump-lime, CaO. 


Hydrated-lime, 
Ca(0H)2. 


Sodium carbonate, Na2C03 

Magnesium chloride, MgCl^ 

Magnesium sulphate, MgS04 

Magnesium bicarbonate, Mg (HC03)o 


0.529 
0.589 
0.466 
0.767 
1.330 
0.346 
0.560 


0,699 
0.778 
0.616 
1.014 


Magnesium carbonate, MgCOs 

Calcium bicarbonate, Ca (HC03)2 

Calcium carbonate, CaCOs 


1.757 
0.457 
0.740 







If any of the salts tabulated above occur in a water analysis multiply 
the amount of each by the corresponding factor and the product will 
represent the weight of reagent to be used. 

The sulphates and chlorides of sodium and potassium in raw water 
need not be considered since they do not add to the hardness. 

If a water has been properly softened there will be little if any scale 
since the small amount of lime and magnesia salts left in the water are 
of such a character that when precipitated as a result of concentration 
in the boiler only a shght sludge is formed. This sludge can be kept 
at a minimum by proper blowing off. 

The lime-soda process does not eliminate all the scale-forming salts 
but removes a large part of them. The precipitates formed are them- 



TABLE 102. 

EFFECT OF SODA-LIME TREATMENT AND FILTRATION. 
Niagara River — Buffalo, N. Y. 



Raw. 


Gallon. 


Treated. 


Gr. per 

U.S. 

Gallon. 


Volatile and organic matter . . 

Silica 

Oxides of iron and alumina. . . 

Calcium carbonate 

Calcium sulphate 

Magnesium carbonate 

Magnesium chloride 


trace 
1.85 

trace 
2.20 
2.11 
0.48 
0.05 
1.16 
0.76 
8.61 
0.10 
1.43 
7.85 


Volatile and organic matter . 

Silica 

Oxides of iron and alumina. . 

Calcium carbonate 

Magnesium hydrate 

Sodium sulphate 

Sodium chloride 


trace 
0.15 

trace 
1.25 
0.25 
2.21 
80 


Magnesium nitrate. . . 


Sodium nitrate. . . 


1 31 


Sodium chloride 


Total solids 




Total solids 


5 97 


Suspended matter 


Incrusting substances 




Free carbonic acid 

Incrusting substances 


1 65 







Cost of treatment, 0.8 cent per 1000 gallons. 



FEED WATER PURIFIERS AND HEATERS 577 

selves partly soluble in water and it is therefore impossible to reduce 
the hardness below, say, 4 grains per gallon. 

Table 102 shows the influence of soda-lime treatment in a specific 
case. 

Causticity, as used in water treatment, is a term to indicate the 
presence of an excess of lime added during treatment. Alkalinity is a 
general term used for the presence of compounds having the power to 
neutralize acids. For an excellent discussion of this subject consult 
"Water for Steam Boilers — Its Significance and Treatment" by 
Scott & Bailey, Jour. A.S.M.E., Nov., 1916, p. 867. 

Caustic Soda and Boiler Corrosion: Prac. Engr., Feb. 15, 1916, p. 211. 

267. Water-softening and Purifying Plants. — The term "water-soften- 
ing" is ordinarily applied to systems in which the temporary and per- 
manent hardness of the water are eliminated or reduced to a minimum, 
whereas the term "purifying" refers to systems in which some par- 
ticular impurity or impurities are neutralized or completely removed. 
In boiler practice these terms are used synonymously and are applied 
to all systems of water treatment outside the boiler. Water-softening 
plants include two types of cold processes, the intermittent and the 
continuous; and the hot process. The cold-process plant is used chiefly 
in softening waters for locomotives and in large plants where water is 
used in considerable quantities. The hot process is commonly used in 
plants where exhaust steam is available for heating the water. 

A typical continuous system is illustrated in Fig. 334. The hard 
water enters the softener through the inlet pipe, is discharged into the 
raw water box, whence it passes over the water wheel, and thus generates 
the power necessary to maintain the reagents in constant agitation. 
From the water wheel the hard water passes into the top of the cone, 
where it meets the reagents delivered by the lift pipe and is thoroughly 
mixed with them. The reagents are dissolved in the mixing tank, lo- 
cated at the ground level, and by means of a steam, electric, or power 
pump are then elevated into the chemical tank above. One charge is 
sufficient to last ten hours or more. The reagents are apportioned to 
the amount of incoming raw water to the dividing box. (Inasmuch 
as the "head" over this stream varies directly with any fluctuation of 
the main hard water stream, the two streams are constantly maintained 
in the same proportion to each other.) In the dividing box this small 
stream is again divided by a slide which throws one part of the water 
back into the hard water stream and another part — which determines 
the rate of flow of the chemicals — into the regulating tank. As the 
level of water in the regulating tank rises, the float rises Hkewise and 



578 



STEAM POWER PLAxMT ENGINEERING 




"Sludge Sump 

Fig. 334. Kennicott Type K Feed-water Purifier. 



FEED WATER PURIFIERS AND HEATERS 



579 



by means of a connecting chain lowers the mouth of the hft pipe in the 
reagent tank. Through this Uft pipe the reagents flow into the top 
of the cone and intimately mix with the raw water. The reaction 
between the raw water and the reagents starts as soon as they meet, 
and as the mixture flows from the mixing plate into the reaction cone 
or downtake, the precipitation of the scale-forming and soap-destroying 
material commences to take place. Flowing at a constantly decreasing 
rate, owing to the constantly increasing diameter of the channel, the 
water passes to the bottom of the cone, turns and flows upward still at 
a constantly decreasing rate, the precipitate falhng away from it as it 
moves. Finally the water passes through a filter which removes any 
slight trace of precipitate that remains; and it then is discharged from 
the top of the softener. The precipitate, which consists of the impuri- 
ties of the raw water and the softening chemicals in chemical union, 
falls to the bottom of the main tank and is from time to time discharged 
therefrom through a sludge valve. An electric indicator is provided 
which rings a bell half an hour before a new supply of reagents is needed 
and thus notifies the attendant of the fact. The lift pipe is a tube, 
flexible for a portion of its length, through which the chemicals leave 
the chemical tank. By means of the regulating device the mouth of 
this tube is maintained at a constant depth of immersion in the surface 
of the dissolved reagents. 

In the Scaife system for water purification feed water first enters the 
heater, where it attains a temperature of from 200 to 210 deg. fahr. 
As a portion of the free CO2 is driven off by the heat the carbonates of 
lime and magnesia are precipitated and are deposited in removable pans 



PRECIPITATING 
TANKS, 



SDLUTinN TANKS 




LJi—jl i L.jL.ji iL.j LJ Lj LJ Li l— i 

Fig. 335. General Arrangement of Scaife System of Feed-water Purification. 

inside the heater. On its way the heated water is forced by the boiler 
feed pump into a large precipitating tank, where the necessary chemicals 
are introduced by two small pumps. These pumps take the solution 
of chemicals from the solution tanks which hold a sufficient quantity 
to operate the plant from eight to twelve hours. The precipitating tank 
is so constructed as to cause intimate and thorough mixing of the 
chemicals with the water. Thus the acids are neutraUzed, and the 



580 



STEAM POWER PLANT ENGINEERING 



scale-forming substances are precipitated by being changed to insoluble 
substances which sink to the bottom of the precipitating tank, whence 
they are readily removed. Some of the hghter substances remaining 
in suspension are carried along with the water as it passes into the 
filters, which effectively remove all suspended matter. This system 
is continuous in operation, and purification is accomplished without 
appreciably retarding the onward flow of feed water. Fig. 335 shows 



XH 


Px 


[^>Engihe 


^^ 


r\ 


^Treating 


Tank X 


Inlet :i f 


Treating 


TankX^ 


^ 


© 




^ 




a » 1 top 




% 


<^ 




^^ 


1 


V. 




J 


|i\\ 






_y 


J" 


w \ 


i 




Outlet 1/ 


Filter Jj 








Washout \ 


K^ 











•^ 


r* F 


^ 


, Clutch 


fc-^^ 


»- 




O Float 

V 

^Stirrinr 


i d — 

Device%» 


1 


















































Tq^ " _ 






A 


^-X-pm - itttn 


J 


n n 4 


\^ t J\. lY 


W M M II M W 



Fig. 336. General Arrangement of We-Fu-Go System of Feed-water Purification. 

a modification of the system." The chemicals are pumped from the 
''chemical tank" into the ''solution tanks," where the feed water and 
chemical solution are thoroughly mixed. The treated water is taken 
from these tanks and pumped into the "precipitating tanks" where a 
large portion of the scale-forming element is precipitated. From the 
precipitating tanks the water is forced through a series of filters to 
the boiler. 



FEED WATER PURIFIERS AND HEATERS 



581 



Fig. 336 illustrates the We-Fu-Go system of water purification. In 
this installation the water supply first enters the setthng or treating 
tanks into which the chemicals are fed. A thorough mixture is effected 
by the use of the two armed paddles located near the bottom of the 
tanks. From the treating tanks the water flows by gravity into the 
filters, which remove all remaining impure solid matter which does not 
settle to the bottom of the treating tank. The pipes conducting the 
water from the settling tanks to the filter are fitted with a flexible joint 
and float so that the outlets are near the surface at all times, rising and 
faUing with the water level. From the filters the purified water gravi- 
tates into the clear water storage reservoir, from which it is pumped 
into an open heater and thence to the boiler. This system is intermit- 
tent in operation, and in order to provide sufficient time for thorough 




Fig. 337. Anderson System for Preventing Corrosion in Condensers. 

chemical treatment of large quantities, two or more setthng tanks are 
employed. Both the We-Fu-Go and Scaife systems are modified in a 
number of ways to meet different conditions. 

Fig. 337 shows the general arrangement of the Anderson system 
for preventing corrosion in condensers and removing oil from condensed 
steam. The method consists in injecting into the exhaust steam as it 
passes from the preheater to the condenser a solution containing a 
coagulant which changes the emulsion of the cyhnder oil to a flaky 
condition so that it may be separated by setthng, flotation, or filtering. 
The air pump delivers the water to the settling tank F, whence it is 
taken to the open gravity filters G, G, of a superficial area proportional 
to the amount of water to be passed and containing a filter bed of four 
feet of crushed quartz. This will run about four days without any 
marked difference in efficiency, after which time the bed is stirred to a 
depth of two feet by mechanical agitators and flushed with clean water, 
by which all impurities are carried to the sewer. The solution is pre- 



582 STEAM POWER PLANT ENGINEERING 

pared in tank A, in which the water level is preserved by a ball float 
and into which filtered water is admitted through pipe B, while the 
substance with which the water is treated is pumped in through the 
pipe D by a small pump operated from the main engine. The flow to 
the ''rose head" above the condenser is controlled by the valve E, 
and a meter in this pipe records the amount being fed. The water 
ordinarily required for ''make up" is sufficient to carry in the solution. 
There is very little loss of water, and the rapid corrosion of the con- 
denser tubes, which has been so great an obstacle to the successful use 
of surface condensers, is much reduced. The chemicals used perform 
a twofold duty, viz., to neutralize the water and make it chemically 
inactive and to coagulate the oily matter contained in the steam so that 
mechanical filtration is possible. (Power, June, 1903, p. 304.) 

Fig. 338 shows a side elevation and a sectional end elevation of a 
Permutit water-softening plant. Referring to the sectional end eleva- 
tion it will be seen that the raw water is delivered to the top of a closed 
tank and is caused to percolate successively through a layer of crushed 
marble, Permutit and gravel. This filtration effects the necessary 
purification. Permutit is the trade name of artificially manufactured 
hydrous silicates produced from clay, feldspar, soda ash and pearl ash. 
Permutit has the property of automatically eliminating all hardness 
from the water passing through it. It is a process of exchange, the 
calcium and magnesium in the water being replaced by the sodium in 
the Permutit. The softening continues until the sodium is used up. 
After the latter is exhausted the Permutit may be restored to its origi- 
nal efficiency by soaking it in common brine. The calcium and mag- 
nesium are thrown off and the sodium from the brine takes their place. 
Permutit is insoluble and a large excess of the reagent may be used 
without producing causticity, since it automatically gives up only 
enough soda to effect the required softening. See Power, Feb. 8, 
1916, p. 198 and "Chemistry of Permutit," pamphlet published by 
the Permutit Company, N. Y. 

Water-softening plants cost from $4 to $5 per horsepower for plants 
of 1000 horsepower and less, from $3 to $4 for plants of 1000 to 2000 
horsepower, and as low as $1.50 for plants of 5000 horsepower or more. 
The depreciation of wooden tanks is as high as 15 per cent a year, while 
that of steel tanks should not be greater than 5 per cent. Unless wooden 
tanks are considerably cheaper than steel tanks they are not a good 
investment. The cost of water purification varies from a fraction of a 
cent to 2 cents per 1000 gallons, depending upon the size of the plant 
and the quantity and character of the impurities. (American Elec- 
trician, March, 1905, p. 125.) 



p 



FEED WATER PURIFIERS AND HEATERS 



583 




i 



- ^;^^M|f»«^^^^ 




c3 
faC 



m 



584 STEAM POWER PLANT ENGINEERING 

Water Softening and Treatment for Power Plant Purposes: Chem. Engr., Jan., 
1910, p. 5; Eng. News, June 6, 1912, p. 1087; Ry. Age Gazette, Aug. 16, 1912, 
p. 288; Ry. Master Mechanic, May, 1910, p. 153; Power, May 28, 1912, p. 780; 
Apr. 18, 1911, p. 598; Prac. Engr., U. S., Mar., 1910; see (Serial) 1915. 

368. Economy of Preheating Feed Water. — Although a feed-water 
heater acts to some extent as a purifier its primary function is that 
of heating the feed water. Generally speaking, for every 10 degrees 
that the feed water is heated there is a gain in heat of 1 per cent and a 
corresponding saving of coal, if the heat which warms the feed water 
would otherwise be wasted. Again, the smaller the difference in tem- 
perature between the steam and the feed water the less will be the 
strain on the boiler shell due to unequal expansion and contraction, an 
item of no small consequence. 

If H represents the heat content of the steam above 32 deg. fahr., 
to the temperature of the cold water, and t the temperature of the water 
leaving the heater, then *S, the per cent gain in heat due to heating the 
feed water, may be expressed 

'-'^H-^0Wy (236) 

The expression is not theoretically correct, since it assumes a con- 
stant value of unity for the specific heat, whereas the specific heat varies 
with the temperature. The variation is so slight, however, that it 
may be neglected for all practical purposes. 

Example 51. Steam pressure 100 pounds gauge; temperature of 
water entering heater 80 deg. fahr. ; temperature of water leaving heater 
210 deg. fahr. Required, saving due to heating the feed water. 

Here H (from steam tables) is 1188, U = SO, t = 210. 

5 = 100 (210-80) 



1188 - (80-32) 
= 11.4 per cent. 

This equation gives the thermal saving only, and the first cost of the 
heater, interest, depreciation, attendance, and repairs must be taken 
into consideration before the net saving measured in dollars and cents 
is ascertained. In the average installation the net saving is a substan- 
tial one. 

Table 103 based upon equation (236) may be used in determining the 
percentages of saving due to the increase in feed-water temperature. 

Feed-water Heating. — Fower, June 25, 1912; Eng. News, Sept. 9, 1909, p. 284; 
Elec. Wld., March 2, 1911, p. 551; Mech. Engr., Nov. 5, 1909, p. 588; Engr. U. S., 
Jan. 1, 1906, p. 8, Aug. 15, 1904, p. 15; St. Ry. Jour., July 22, 1905, p. 145. 



FEED WATER PURIFIERS AND HEATERS 



585 



TABLE 103. 

PERCENTAGE OF SAVING FOR EACH DEGREE OF INCREASE IN TEMPERATURE 

OF FEED WATER. 
(Based on Marks & Davis Steam Tables.) 



Initial 








Boiler Pressure Above Atmosphe 


re. 








Temp, 
of Feed. 





20 


40 


60 


80 


100 


120 


140 


160 


180 


200 


32 


.0869 


.0857 


.0851 


.0846 


.0843 


.0841 


.0839 


.0837 


.0835 


.0834 


.0834 


40 


.0875 


.0863 


.0856 


.0853 


.0849 


.0846 


.0845 


.0843 


.0841 


.0840 


.0839 


50 


.0883 


.0871 


.0864 


.0859 


.0856 


.0853 


.0852 


.0850 


.0848 


.0847 


.0846 


60 


.0891 


.0878 


.0871 


.0867 


.0864 


.0861 


.0859 


.0857 


.0855 


.0854 


.0853 


70 


.0899 


.0886 


.0879 


.0874 


.0871 


.0868 


.0867 


.0865 


.0863 


.0862 


.0861 


80 


.0907 


.0894 


.0887 


.0882 


.0878 


.0876 


.0874 


.0872 


.0871 


.0870 


.0869 


90 


.0915 


.0902 


.0895 


.0890 


.0887 


.0884 


.0882 


.0880 


.0878 


.0877 


.0876 


100 


.0924 


.0910 


.0903 


.0898 


.0895 


.0892 


.0890 


.0888 


.0886 


.0885 


.0884 


110 


.0932 


.0919 


.0911 


.0906 


.0903 


.0900 


.0898 


.0896 


.0894 


.0893 


.0892 


120 


.0941 


.0927 


.0919 


.0915 


.0911 


.0908 


.0906 


.0904 


.0902 


.0901 


.0900 


130 


.0950 


.0936 


.0928 


.0923 


.0919 


.0916 


.0915 


.0912 


.0911 


.0910 


.0909 


140 


.0959 


.0945 


.0937 


.0931 


.0928 


.0925 


.0923 


.0921 


.0919 


.0918 


.0917 


150 


.0969 


.0954 


.0946 


.0940 


.0987 


.0933 


.0931 


0930 


.0928 


.0927 


.0926 


160 


.0978 


.0963 


.0955 


.0948 


.0946 


.0942 


.0940 


.0838 


.0936 


.0935 


.0934 


170 


.0988 


.0972 


.0964 


.0958 


.0955 


.0951 


.0948 


.0947 


.0945 


.0944 


.0943 


180 


.0998 


.0982 


.0973 


.0968 


.0964 


.0960 


.0958 


.0956 


.0954 


.0953 


.0952 


190 


. 1008 


.0992 


.0983 


.0977 


.0973 


.0969 


.0968 


.0965 


.0964 


.0963 


.0962 


200 


.1018 


.1002 


.0993 


.0987 


.0983 


.0978 


.0977 


.0974 


.0973 


.0972 


.0971 


210 


.1029 


.1012 


.1003 


.0997 


.0993 


.0989 


.0987 


.0984 


.0983 


.0982 


.0981 


220 




.1022 


.1013 


.1007 


.1003 


.0999 


.0997 


.0994 


.0992 


.0991 


.0990 


230 




.1032 


.1023 


.1017 


.1013 


.1009 


.1007 


.1004 


.1002 


.1001 


.1000 


240 




. 1043 


.1034 


.1027 


.1023 


.1019 


.1017 


.1014 


.1012 


.1011 


.1010 


250 




.1054 


.1044 


.1008 


.1034 


.1029 


.1027 


.1024 


.1022 


.1021 


.1020 



Multiply the factor in the table corresponding to any given initial temperature of feed water and boiler 
pressure by the total rise in feed-water temperature; the product will be the percentage of saving. 



269. Classification of Feed-water Heaters. — Feed-water heaters may 
be classified according to the source of heat, as 

1. Exhaust steam, in which the heat is received from the exhaust of 
engines, pumps, etc. 

2. Flue gas, in which the waste chimney gases are the source of the 
heat. 

3. Live steam purifiers, or those using steam at boiler pressures; or 
according to the method of heat transmission, as 

1. Open heaters, in which the steam and feed water mingle and the 
steam in condensing gives up its heat directly to the water. 

2. Closed heaters, in which the steam and water are in separate 
chambers and the steam gives up its heat to the water by conduction. 

Heaters may also be classified according to the pressure of the heat- 
ing steam, as 

1. Vacuum or primary, in which the pressure is less than atmospheric 
and applies particularly to heaters utihzing the exhaust of condensing 
engines. These are always of the closed type. Open heaters in which 
the pressure is less than atmospheric are not usually classed as vacuum 



586 STEAM POWER PLANT ENGINEERING 

heaters. 2. Atmospheric or secondary, in which the pressure is atmos- 
pheric or, Hterally, that corresponding to the back pressure on the 
engines and pumps. 

3. Pressure, in which the pressure corresponds to that in the boiler 
and in which the heat is used primarily for purifying purposes. 

CLASSIFICATION OF A FEW TYPICAL HEATERS. 

[ Cochrane 
Open . . .Atmospheric .S?i?i!!n 



Exhaust steam 



StiUwell 
i Webster 

f Wain Wright I Water 
Wheeler . . . . ( Tube 
Otis i Steam 



pressure | Otis i Steam 

Berryman . . ) Tube 
Green 

Flue Gas •! American 

1 Sturtevant 

Live Steam Open Pressure | g°P'^''^a„^th 

Heaters may be still further classified as 

1. Induced, in which only such steam is admitted as is induced by 
its condensation. That is, the feed water condenses the steam. This 
creates a partial vacuum which draws in more steam. 

2. Through, in which all the steam is forced through the heater 
irrespective of condensation. 

270. Open Heaters. — Fig. 339 gives a sectional view of a Cochrane 
special feed heater and receiver and is a typical example of an open 
heater. Exhaust steam enters the heater through a fluted oil separa- 
tor as indicated, and passes out at the top, while the oily drips are 
automatically drained to waste by a suitable ventilated float. The 
feed water enters through an automatic valve and is distributed over 
a series of copper trays so arranged and constructed that the water is 
forced to fall in a finely divided stream before reaching the reservoir in 
the bottom. The steam coming in contact with the water particles 
gives up latent heat and condenses. Some of the scale-forming element 
is deposited on the surface of the trays, from which it may be removed. 
The suspended matter is eliminated by a coke filter in the bottom of 
the chamber, and the floating impurities are decanted by a skimmer 
or overflow weir. The particular heater shown in the illustration is 
especially designed for use in a steam-heating plant; i.e., besides per- 
forming all the functions of an open heater, it provides for the reception 
and heating of the condensation returned to it from the heating system. 

Fig. 340 shows a section through a Hoppes open heater, illustrating 
the "pan" type. Exhaust steam enters at H, passes through oil filter 0, 
and completely surround pans T, T. The feed water enters at B, 



FEED WATER PURIFIERS AND HEATERS 



587 




Fig. 339. Cochrane Feed- water Heater. 




Fig. 340. Hoppes Horizontal Feed-water Heater. 



588 



STEAM POWER PLANT ENGINEERING 



and the rate of flow is regulated by valve F, which is controlled by a 
suitable float in the lower part of the chamber. The water in flowing 
over the sides and bottoms of the pans comes in direct contact with 
the steam. 

271. Combined Open Heater and Chemical Purifier. — Combined 
feed-water heaters and chemical purifiers are finding increased favor 
with some engineers in many districts where the feed water is particu- 
larly bad and when space limitations preclude the use of water-softening 
plants. Although better than the plain open heater the purification is 
not thorough because of the short time that the water is in the heater. 

372. Temperatures in Open Heaters. — The temperature to which feed 
water is raised in an open heater may be determined as follows: 

Let H represent the heat content of the steam entering the heater, 
^0 the temperature of the water entering heater, 
t the temperature of the water leaving heater, and 
S the ratio of exhaust steam to the feed water, by weight. 

Then, allowing a loss of 10 per cent due to radiation, etc., 0.9 S 
(H — t -\- 32) will be the B.t.u. given up by the exhaust steam to each 
pound of feed water, and {t — U) will be the B.t.u. absorbed by each 
pound of water. 

Therefore 0.9 S {H - t + d2) = t - U, from which 



t = 



to + 0.9 S(H + 32) 
H-0.9>S 



(237) 



TABLE 104. 

FINAL FEED-WATER TEMPERATURES. OPEN HEATER. 
(Temperature of steam, 212 degrees F.) 









Initial Temperature of Feed Water 


, Degrees F. 






40 


50 


60 


70 


80 


90 


100 


110 


120 


130 


1 


2 


60.1 


69.9 


79.7 


89.5 


94.4 


109.2 


119.0 


128.8 


138.7 


148,5 


p 


3 


69.9 


79.6 


89.3 


90.1 


108.8 


118.6 


128.3 


138.0 


147.8 


157.5 


s 


4 


79.5 


89.1 


98.8 


108.5 


118.1 


127.8 


137.4 


147.1 


156.7 


166.4 


i^ 


5 


89.0 


98.5 


108.1 


117.7 


127.2 


136.8 


146.4 


155.9 


165.5 


175.1 


'^'u 


6 


98.3 


107.7 


117.2 


126.7 


136.2 


145.7 


155.2 


164.7 


174.2 


183.6 


-•ri 


7 


107.4 


116.8 


126.2 


135.6 


145.0 


154.4 


163.8 


173.2 


182.5 


192.1 


^i 


8 


116.4 


125.7 


135.0 


144.4 


153.7 


163.0 


172.4 


181.8 


191.0 


200.3 


o^, 


9 


125.2 


134.5 


143.7 


153.0 


162.2 


171.5 


180.7 


190.0 


199.2 


208.5 


"S-^ 


10 


133.3 


143.1 


152.3 


161.4 


170.6 


179.8 


189.0 


198.1 


207.3 


212.0 


o 


11 


142.5 


151.6 


160.7 


169.7 


178.9 


188.2 


197.0 


206.2 


212.0* 


212.0* 


g 


12 


150.9 


159.9 


168.9 


177.9 


187.0 


196.0 


205.0 


212.0* 


212.0* 


212.0* 



* All of the steam not condensed. 



FEED WATER PURIFIERS AND HEATERS 589 

If more steam passes through the heater than can be condensed by 
the feed water, then this equation gives t a fictitious value; in other 
words, t can never be greater than the temperature of the exhaust steam. 

Substituting t = 212, the maximum obtainable temperature with 
exhaust steam at atmospheric pressure, and solving for *S, we find that 
only 17 per cent of the main engine exhaust is necessary to heat the 
feed water to a maximum, tp is assumed to be 60 deg. fahr. 

Table 104 has been determined from this equation and gives the final 
temperatures obtainable in open heaters for various conditions of 
operation. 

Example 52. A power plant has 1200 i.hp. of engines using 20 
pounds of steam per i.hp-hour. AuxiUaries use 2400 lb. steam per hr. 
Pressure in heater pounds gauge, temperature of hot-well supply 
110 deg. fahr. Required temperature of feed water leaving heater. 

Here H = 1150 (from steam tables), ^o = HO, >S = 0.10. 

Substituting these values in (237), 

0.9 X 0.10 (1150 -t-\-32)=t- 110. 

t = 198 deg. fahr. 

213. Pan Surface Required in Open Feed-water Heaters. — Pan or 

tray surface required varies according to the quality of the water with 
regard to both scale-making material and mud, and may be approxi- 
mated by the formula 

Pan surface, sq. ft. = Pounds of water heated per hour ^^ss) 



Horizontal 
Type. 




For very muddy water, c 118 110 

Slightly muddy water, c 166 155 

For clean water, c 500 400 



374. Size of Sliell, Open Heaters. — General proportions of open heaters 
vary considerably on account of the different arrangements of pans or 
trays, filter and oil-extracting devices. A fair idea of the size of shell 
required may be obtained by the formulas 



. r 1 11 Horsepower 

Area oi shell = -, n— : — -^ — , 

a X length in feet 



(239) 



Length of shell = ^^^^^P^^^^ , , (240) 

a X area in square feet 

a = 2.15 for very muddy water, 
a = 6 for slightly muddy water, 
a = S for clean water. 



590 STEAM POWER PLANT ENGINEERING 

The horsepower in this case is obtained by dividing the weight of water 
heated per hour by the steam consumption of the engine per horse- 
power per hour. 

Pans containing 2.5 square feet and less are usually made round, and 
larger sizes rectangular in plan. When circumstances will permit it 
is better to have not more than six pans in any one tier, since it is 
advisable to proportion the pans so as to obtain as low a velocity over 
each as practicable. 

Distance between trays or pans is seldom less than one-tenth the 
width for rectangular and one-fourth the diameter for round pans. 
Volume of storage and settling chamber in horizontal heaters varies 
from 0.25 for good quality of water to 0.4 of the volume of the shell 
for muddy water, 0.33 being about the average. In the vertical type 
the setthng chamber represents respectively 0.4 and 0.6 the volume of 
the shell with clear and muddy water. Filters occupy from 10 to 15 
per cent of the volume of the shell in the horizontal type and from 15 
to 20 per cent in the vertical type, the smaller percentage correspond- 
ing to clear water and the larger to muddy water or water containing a 
considerable quantity of impurities. 

Open Heaters: Cassier's Mag., Aug., 1903, p. 33; Engr. U, S., Jan. 1, 1906, pp. 
17, 78; St. Ry. Jour., Feb. 4, 1905, p. 227; Elec. Wld., Apr. 27, 1911, p. 1051. 

375. Types of Closed Heaters. — Closed heaters may be grouped into two 

classes : 

1. Water tube, Fig. 341, and 

2. Steam tube, Fig. 345. 

Closed heaters, both water tube and steam tube may operate with 

1. Parallel currents, where the water and steam flow in the same 
direction, Fig. 344, or with 

2. Counter currents, where the water and steam flow in opposite 
directions. Fig. 343. 

Water-tube heaters may be still further classified as 

1. Single-flow, in which the water flows through the heaters in one 
direction only. Fig. 341. 

2. Multi-flow, in which the water flows back and forth a number of 
times, as in Fig. 343. 

3. Coil heater, in which the water flows through one or more coils, 
as in Fig. 344. 

4. Film, in which the water is forced across the heating surface in a 
thin sheet or film. 

276. Water-tube Closed Heaters. — Fig. 341 shows a section through 
a feed-water heater of the single-flow straight-tube type. The tubes 



FEED WATER PURIFIERS AND HEATERS 



591 



Surface Blow 



are of plain brass and the shell of cast iron. The tubes are expanded 
into the tube sheets by a roller expander. To provide for expansion 
the upper tube sheet and water chamber are secured to the main shell 
by means of a special exnansion joint the details of which are shown 

in Fig. 342. R is sl ring or 
gasket of soft annealed copper 
and G, G two gaskets of special 
packing with brass wire cloth 
insertion. These gaskets form 
a flexible expansion joint be- 
tween C and tube sheet D, so 
that the whole upper chamber, 
which is carried solely by the 
tubes, is free to move up and 
down as the tubes expand or 
contract under varying tem- 
peratures. 



Water 



Drip 




Exhaust from 
Heater 




Fig. 341. Goubert Single-flow 
Closed Heater. 



Fig. 342. Details of Expansion Joint, 
Goubert Heater. 



Fig. 343 shows a section through a Wainwright heater, illustrating 
the multi-flow water-tube type. The body of the heater is of cast iron, 
the tubes of corrugated copper. The water passes through the tubes 
and the steam surrounds them. The feed water and exhaust steam 
do not mingle, and hence the oil in the exhaust does not contami- 
nate the water. The water chambers are divided into several com- 
partments, as shown in the illustration, and the partitions are so 
arranged that the flow of feed water is directed back and forth through 
the various groups of tubes in succession. This arrangement gives a 
higher velocity of flow than the non-return type of heater, and therefore 
increases the rate of heat absorption. The mud and impurities settle 



592 



STEAM POWER PLANT ENGINEERING 



at the bottom and are discharged through the mud blow-off. Such 
impurities as rise to the surface are removed by the surface blow-off. 
The tubes are corrugated to allow for expansion and at the same time 
to increase the transmission of heat. Referring to Fig. 343: Exhaust 
steam enters at A and leaves at E, and the portion which is condensed 




1 Exhaust 

r->nln n n n r^r-in Qutlet 






Fig. 343. Wainwright Multiflow 
Closed Heater. 



Fig. 344. Typical Coil Heater. 



is drawn off at D. Feed water enters at I and is discharged at 0. P, 
P are mud blow-offs and aS is an opening for a safety valve. Fig. 356 
gives results of tests showing the relative efficiencies of plain and corru- 
gated tubes for various velocities. 

Fig. 344 shows a partial section through a Harrisburg feed-water 
heater. This apparatus is a typical example of the coiled-tube heater. 
Three sets of concentric copper coils are brazed to gun-metal manifolds 



FEED WATER PURIFIERS AND HEATERS 



593 



and supported by clamp stays as indicated in the illustration. Feed 
water enters the heater at the bottom manifold and passes through the 
coils to the feed outlet. The exhaust steam enters the heater at the 
bottom and surrounds the coils in its passage to the outlet at the top. 
The coils are designed to withstand a pressure of 600 pounds per square 
inch. 

277. Steam-tube Closed Heaters. — Fig. 345 shows a section through 
an Otis heater, illustrating the steam-tube type. Here the exhaust 
A C 






m^mMn 



Fig. 345. Otis Steam-tube Feed- 
water Heater. 



Fig. 346. Baragwanath Steam-jacketed 
Feed- water Heater. 



steam passes through the tubes which are surrounded by the feed water. 
The exhaust steam enters at A, and passes down one section of tubes 
into the enlarged space of the water and oil separator 0, in w^hich the 
condensation and oil are deposited. From this chamber the steam 
passes up through the other section of tubes to outlet C, thus passing 



594 STEAM POWER PLANT ENGINEERING 

twice through the entire length of the heater. The water enters at E 
and is discharged at G. R is the blow-off opening. The tubes are of 
seamless brass and are curved to allow for expansion. Condensed 
steam is withdrawn at P. 

Fig. 346 shows a partial section through a Baragwanath steam- 
jacketed steam-tube heater. Exhaust steam enters at A, passes up 
through the tubes, returns down annular space E between the inner 
shell and jacket, and passes out at B. Feed water enters at C and leaves 
at D. E is the scum blow-off, G the heater drain, and H the jacket drain. 

378. FUm Heaters. — The heating element in a film heater con- 
sists usually of two spirally corrugated tubes, one within the other, the 
water path being the small annular clearances between the two. Thus 
the water is directed in a spiral path due to the corrugations, and for 
a given velocity the particles of water come more often in contact 
with the heating surface than in plain tubes because they are contained 
within an annular space whose perimeter is large in comparison with its 
area. This type of heater though highly efficient in heat transmission 
necessitates the use of comparatively pure water and is not commonly 
used for feed Avater heating. 

279. Heat Transmission in Closed Heaters. — Since the closed heater is 
practically the same in principle as a surface condenser the laws of 
heat transmission are practically identical in both cases. The tempera- 
ture of the steam and water are higher in the atmospheric heater but 
otherwise the heat exchange is the same in all heaters and condensers 
of the water-tube type. Increasing the velocity of the water passing 
through the heater increases the rate of heat transmission and thereby 
renders the heating surface more effective. In order to employ moder- 
ately high velocities and at the same time allow sufficient time in which 
to raise the temperature to a maximum, the tubes should be as long as 
practicable and of small diameters. Other things being equal, a heater 
containing a large number of tubes of small diameter is more efficient 
than one containing a small number of large tubes. It is important to 
proportion the heater according to the amount of water to be heated 
and the maximum temperature to which the water must be raised. In 
designing a heater, then, the maximum temperature to which the water 
is to be raised and the coefficient of heat transfer are assumed and the 
amount of heating surface is calculated from equations 241 or 242. 

Although recent experiment* shows that the amount of heat trans- 
mitted through the heating surface is proportional to some power of I 
the mean temperature difference the value of the exponent is not fan 
from unity (0.8 to 0.9) and it may be safely taken as such, particularly. 

* Jour. A.S.M.E., Aug., 1915, p. 483. 



FEED WATER PUIUEIERS AxND HEATERS 



595 



in view of the liberal factor allowed in the assumed value of the coeffi- 
cient of heat transfer, U. With this assumption the extent of heating 
surface may be calculated from the following adaptation of equation 
(210) 

cio{k-U)^ (241) 



S = 



Ud 



in which 

S = total tube heating surface, sq. ft., 
c = mean specific heat of water; this may be taken as 1.0, 
w = weight of water heated per hr., 
t2 = final temperature of the feed water, deg. fahr., 
^0 = initial temperature of the feed water, deg. fahr., 
U = mean coefficient of heat transfer for the entire surface, B.t.u. 
per sq. ft. per deg. difference in temperature per hour., 
mean temperature difference between the steam and that of the 
water. 

t2 - to 



d 



d = 



log. 



ts 



(See Equation (219)) 



ts -t2 



Substituting this value of d in equation (241) (taking c = 1) and re- 
ducing we have 

■ ^'log'-^«- (242) 



U 



ta ^5 



For a given extent of heating surface S, the temperature difference 
between that of the steam and the feed water leaving the heater may be 
calculated by solving equation (242) for ts — (2, thus 



ts -to 



(243) 



in which e = base of the Naperian logarithm = 2.718 



n = 



SU 



w 



■^ 


' n 






■"" 


- 


~ 
















1 


3 




























1 


1 










j»X 


et 


^— 


^ 






V 


. 1 






pee 


AJ 


^ 














1 


p. 


" 1 




^ 




-^ 


















1 


1 


1^ 


























1 






























1 


J 


LliJ 


_ 


__ 


._ 


L- 




Lj 












__ 


1 

L_li 



By taking different extents of area 
S and solving for the corresponding 
values of ts — ^2 the temperature gradi- 
ent for a given heater may be obtained 
as illustrated in Fig. 347. 

From equation (241) it will be seen 
that extent of heating surface depends 
upon the weight of water to be heated, the temperature of the steam, 
the desired temperature of the feed-water heater and the value of L\ 



Length of Tube 

Fig, 347. Temperature Gradient in 
Feed-water Heater Tube. 



596 



STEAM POWER PLANT ENGINEERING 



Since the extent of heating surfaces increases rapidly as (2 approaches 
ts, and becomes infinity for ^2 = tg, it is desirable to limit (2 to some 

practical figure. An average maxi- 
mum for t2 = ts — 4. 

The coefficient of heat transfer 
varies within wide limits depending 
upon type of heater and the con- 
ditions of operation, and ranges from 
[/ = 150 in steel tube heaters with 
low water velocities to 1000 or more 
in the film type of corrugated brass 
tube heaters with water velocity of 
7 ft. per second. In practice a lib- 
eral factor is allowed for possible 
heat reduction due to the presence 
of air and the accumulation of oil 
scale or other deposit on the tube 
surfaces. 

For steam coils submerged in 
water and from which the condensa- 
tion is withdrawn as rapidly as it is 
formed the value of U in Table 105a appears to give satisfactory results. 

Example 53. Determine the length of f in. (O. D.), yV in. thick 
brass tubes in a closed heater designed to heat water from 60 to 196 deg. 
fahr., steam temperature 212 deg. fahr., water velocity 2 ft. per sec, 
U = 400. 

S = ^ = ^dl = 0.1971 

I = length of tube, ft. 

2 X 3600 X TTcm 7200 X 3.14 X HY X 62.4 _ __ „ , 

w = , , , . . , = ^ ^ ^ _ ^ = 957 lb. per nr. 



1000 


— 


— 








— 




— 


— 


\7 






— 


'"'doo 




















/ 


.^y 






^000 


















/ 




^/ 








A 
















i 


i 


// 










H 














v 


4 


^ 


/ 


/ 




















i.^/ 


,f/ 


/ 








^700 












t7F/f.*; 


y 










rt 












Wiwi 


7 












^ 












/b'A/ 














Q 












^/ 


vA 




y 










^m 










/f. 




^- 


f 






















^A 


/ 


4 


r 
























^/ 


A 


/ 
















1.300 










^/ 


Y 


























/ 




















ffl 

U nnn 






/ 
























^200 






^/ 








s 


ear 


i21 


2D 


>&• 


Fat 


r. 




s 
■^ 100 




/ 


























ejioo 


/ 


/ 




























/ 





























Fig 



50 100 160 200 250 300 

Water Velocity— Ft. per 3Iin. 

348. Coefficient of Heat Transfer. 
(For General Design.) 



144 X 4 



144 X 4 



Substituting these values in equation (241), 

nin^7 957, 212-60 
0.197Z=^ log. 212-33^- 
From which I = 27.3 ft. approx. 

Example 54. A 200-sq. ft. closed heater is rated at 40,000 lb. of 
water per hour, initial temperature, 60 deg. fahr., temperature steam 
212 deg. fahr., U = 300. Required the final temperature of the water. 

From equation (243), 



= e' 



ts — h 

e = 2.718 

^SU ^ 200 X 300 
^ w • 40,000 



1.5, 



whence 
or 



FEED WATER PURIFIERS AND HEATERS 

2^2 - ^^ = 2.718^.^ 



597 



212 - t2 

h = 172.4 deg. fahr. 

TABLE 105. 

HEAT TRANSMISSION IN CLOSED FEED-WATER HEATERS. 
(Based on Commercial Designs.) 



Type of Heater. 


Coefficient of Heat Transfer, U. 


Range. 


Average. 


Single-flow plain brass tubes 


150- 300 
250- 400 
125- 175 
250- 500 
250- 500 
350- 700 
350- 900 
500-1100 


200 


Single-flow corrugated brass tubes 


300 


Sino'le-flow, steel tubes 


150 


*Spiral coils, plain brass tubes 

Multi-flow plain brass tubes 

Multi-flow corrugated brass tubes 


350 
350 
400 


Plain brass tubes with retarders 


450 


Film heater with corrugated tubes 


600 







For small coils and high water velocities these values may be increased 100 per cent. 

TABLE 105a. 
HEAT TRANSFER —SUBMERGED STEAM COILS. 



Mean Temperature 
Difference. 


Coefficient of Heat Transfer, U. 


Iron. 


Brass. 


Copper. 


50 

100 
150 
200 


100 
175 
200 
225 


200 
275 

375 
450 


220 
300 
400 
475 



Example 55. Determine the size of vacuum and atmospheric heaters 
for a condensing plant of 1200 i.hp. Engines use 20 pounds of steam 
per i.hp-hr.; auxiliaries use the equivalent of 10 per cent of the main 
engine steam; vacuum 25 incnes referred to 30-inch barometer; feed 
water, ^o = 50 degrees; temperature of hot well, ^ = 110 degrees; 
coefficient of heat transmission, U = 300 B.t.u. 

Vacuum or Primary Heater. 
Feed water for main engines, 

20 X 1200 = 24,000 pounds per hour. 
Feed water used by auxiliaries, 

10 per cent of 24,000 = 2400 pounds per hour. 
Total feed, 

W = 24,000 + 2400 = 26,400 pounds per hour. 



598 



STEAM POWER PLANT ENGINEERING 



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FEED WATER PURIFIERS AND HEATERS 



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600 STEAM POWER PLANT ENGINEERING 

From equation (242), 

^ 26,400 134 - 50 

300 ^^' 134 - 110 
= 110 square feet. 

On the basis of J square foot of surface per horsepower the rating of 
this heater will be 

110 X 3 = 330 horsepower. 

Atmospheric or Secondary Heater. 

The temperature of the feed water leaving the atmospheric heater, 
equation (237), will be 

^ ^0 + 0.9 S{H + 32) 
1+0.9 5 
where S = 0.10, ^o = HO degrees, H = 1150B.t.u., 

, 110 + 0.9x0.10(1150 + 32) 

^^^"^^ ' = 1+0.9x0.10 

= 198 degrees. 

The required surface is 

. W , ts-to 
^ = U log. jrz-^^ 

where ts = 212, ^o = HO, ^2 = 198, 

, 26,400 , 212 - 110 
whence ^^-30^^^^- 212-198 

= 175 square feet. 
The horsepower rating will be 

175 X 3 = 525. 
280. Open vs. Closed Heaters. — Open and closed heaters have their 
respective advantages and a careful study of the various influencing 
conditions is necessary for an intelligent choice. The following parallel 
comparison brings out a few of the distinguishing features: 

Open Heater. Closed Heater. 

Efficiency. 

With sufficient exhaust steam for heat- The maximum temperature of the feed 
ing, the feed water may reach the water will always be 2 degrees or more 
same temperature as the steam. lower than the temperature of the 

Scale and oil do not affect the heat steam. 

transmission. Scale and oil deposit on the tubes and 

the heat transmission is lowered. 

Pressures. 
It is not ordinarily subjected to much The water pressure is slightly greater 
more than atmospheric pressure. than that in the boiler when placed 

on the pressure side of the pump as is 
customary. 



FEED WATER PURIFIERS AND HEATERS 



601 



Safely. 
Sticking of the back pressure valve may It will safely withstand any pressure 
cause it to "blow up" if provision is likely to occur, 
not made for such an emergency. 

Purification. 
Since the exhaust steam and feed water Oil does not come in contact with the 



feed water. 
Scale is removed with difficulty. 



mingle, provision must be made for 
removing the oil from the steam. 
Scale and other impurities precipitated 
in the heater are readily removed. 

Location. 
Must always be placed above the pump May be placed anywhere on the pressure 
suction and on the suction side. side of the pump. 

Pumps. 
With supply under suction two pumps One cold-water pump is necessary, 
are necessary and one must handle 
hot water. 

Adaptability. 
Particularly adaptable for heating sys- All vacuum or primary heaters are 
terns where it is desired to pipe the necessarily of this type, 
"returns" direct to heater. 

281. "Through" Heaters. — Fig. 349 shows a typical installation of 
a through heater in a non-condensing plant. 




TO. HKATCR 



Fig. 349. Open Heater Connected as a "Through " Heater. Non-condensing Plant. 

It is evident that all the steam must pass through the heater. Now, 
one pound of exhaust steam in condensing gives up approximately 
1000 B.t.u. Hence, if the initial temperature of the feed water is 50 

degrees and the final temperature 210, the engine furnishes ^-^r ~ 



602 



STEAM POWER PLANT ENGINEERING 



aNB 



= 6.26, say, six times the quantity necessary for heating the feed water 
to a maximum. Therefore the area of the pipe supplying the heater 
with steam need be but one sixth that of the main exhaust. With the 
heater connected as in Fig. 349 the connections must 
necessarily be the same size as the exhaust pipe. 

With this arrangement the heater cannot be " cut 
out" while the engine is in operation and hence it 
is not adapted for plants working continuously. 
For the purpose of cutting out a heater while the 
plant is in operation a through heater may be by- 
passed as in Fig. 350. Advantage may be taken 
here of the permissible reduction in the size of 
pipes and fittings, i.e., valves, etc., at C and D 
need be but one half the size of those at A. This 
reduction in size may prove to be a considerable 
Fig. 350. j^^j^ jj^ large installations. 

283. Induced Heaters. — Fig. 351 shows a typical installation of an 
induced heater in a non-condensing plant and Fig. 352 an induced pri- 
mary heater in a condensing plant. 

In the arrangement in Fig. 351 the number of fittings is reduced to a 
minimum and the heater may be readily cut out. Since induced heaters 




Cold'Watcr Supply 




Fig. 351. Open Heater Connected as an "Induced" Heater. Non-condensing Plant. 

are apt to become air-bound, a vapor pipe or vent is inserted in the 
top of the heater as shown. This pipe varies from | to 1| inches in 
diameter, depending upon the size of heater. 

Closed Heaters: Am. Elecn., May, 1900, p. 236, July, 1900, p. 354, Oct., 1905, 
p. 530; Cassier's Mag., Aug., 1903, p. 330; Eng. U. S., Jan. 1, 1906, p. 13; Power, 
April, 1902, p. 11. 



FEED WATER PURIFIERS AND HEATERS 



603 




Fig. 352. Closed Heater Connected as an "Induced" Heater. Condensing Plant. 



Sack Pressure Valve 



-Nou.Return Air Valves 




Fig. 353. Open Heater in Connection with a Low-pressure Turbine. 



604 



STEAM POWER PLANT ENGINEERING 



283. Live-steam Heaters and Puriflers. — The function "of a live- 
steam heater and purifier is primarily that of purification and hence it 
is not ordinarily installed unless the feed water contains scale-forming 
elements such as sulphates of lime and magnesia. These, as previously 




Fig. 354. 



Outflow 

Open Heater in Connection with a Jet Condenser. 



stated, are not entirely precipitated until a temperature of approxi- 
mately 300 deg. fahr. is reached; hence no amount of heating with ex- 
haust steam atmospheric pressure will thoroughly purify feed water 
containing these elements. 

Fig. 355 shows a section through a Hoppes live-steam purifier. Since 
the purifier is subjected to full boiler pressure, the shell and heads are- 
constructed of steel. Within the shell are a number of trough-shaped 
pans or trays placed one above another and supported on steel angle 
ways. Steam from the boiler enters the chamber at A and comes in 
contact with feed water and condenses. The water on entering the 
heater at B is fed into the top pan and, overflowing the edges, follows 
the under side of the pan to the center and drops into the pan below. 
It flows over each successive pan in the same manner until it reaches 
the chamber at the bottom, whence it gravitates to the boiler through 
pipe C. As the steam inclosed in the shell comes in contact with the 
thin film of water, the solids held in solution are separated and adhere 
to the bottom of the pans in the same manner that stalactites form on 
the roofs of natural caves. Authentic tests show that live-steam heaters 
may increase the boiler efficiency. (See Power, Feb. 21, 1911, p. 295.) 



FEED WATER PURIFIERS AND HEATERS 



605 



The purifier should be set in such a position as will bring the bottom of 
the shell two feet or more above the water level of the boilers, as in Fig. 
356. N is the feed pipe from pump to purifier and should be provided 




Fig. 355. Hoppes Live-steam Purifier. 

with a check valve. D is the gravity pipe through which the purified 
water flows to the boiler. This pipe should be carried below the water 
level of the boilers and all branch pipes should be taken off below the 




Typical Installation of a "Live-steam" Purifier. 



water line. Pipe L leads from top of pipe S to pump or other steam- 
using device. This is necessary in order that air and other non-condenfe- 
able gases liberated from the water may be removed from the purifier, 
which would otherwise become air-bound. In the illustration the feed 



606 



STEAM POWER PLANT ENGINEERING 



pump takes its supply from an exhaust steam heater C, The purifier 
is provided with a suitable by-pass so that the water may be fed directly 
to the boiler when necessary. 

Live Steam Heated Feed Water: Elec. Engr., Lond., June 29, 1906; Cassier's Mag., 
Oct., 1911, p. 543; Elec. Rev., Lond., May 20, 1898, p. 667; Eng. Rec, Aug. 30, 
1898, p. 467; Power, March 31, 1908, p. 498, Feb. 21, 1911, p. 295. 

384. Distillation of Make-up Water. — In large central stations equipped 
with turbines and surface condensers the condensate furnishes a supply 
of distilled water for boiler feed purposes. To provide for leakage losses 
an additional supply of water must be had from some other source. 



Head Tank 



Overflow to 
Storage Tank 



House Alternator 
Exhaust - 



Adjustable Back' 
Pressure Valve 



20,000 Kw. Turbine 



Boiler Feed Pump 
Exhaust 




35,000 Sq. Ft. / |— I 
Condenser (32,500 ( 
Sq. Ft. now installed) V "^ 





1000 K\v. House _ 
\lternator (Turbine Boiler Feed 
Driven) Pump (Turbine 

Driven) 



Hot Well Pump 



Barometric 

Injection 

Pump 



Storage ji Tank 



Surge Pump^ 



Fig. 357. Feed-water Heating System at the Connors Creek Station of the Detroit 

Edison Co. 



In some situations raw make-up water is sufficiently pure to warrant 
its introduction into the system without treatment but in most cases 
it is too hard for direct use even though the quantity required is a rela- 
tively small percentage of the total weight of water fed to the boilers. 
In a number of recent installations all make-up water is distilled, thus 
insuring a continuous supply of pure water. Fig. 357 illustrates the 
principles of the feed-water system as installed in the Connors Creek 
Station of the Detroit Edison Company. The condensate from the 
main surface condensers is discharged into one end of a large tank shown 
as the boiler feed tank. A centrifugal pump draws its water from the 
same end of this tank and discharges it into the head of a barometric 
condenser. The relatively cold condensate is picked up by the second 



FEED WATER PURIFIERS AND HEATERS 



607 



pump before it has time to mix with the mass of water in the tank and 
serves as injection water for the barometric condenser. The house- 
service alternator turbine and the boiler feed pump turbine exhaust 
into this barometric condenser so that the condensate from the main 
unit takes up all the heat of the auxiliary steam. The foot of the 
barometric condenser is immersed in the hot end of the boiler feed 
tank. The mixture is then picked up by the boiler feed pump and 
delivered to the boilers. The barometric condenser is therefore the 
equivalent of an open feed-water heater in which exhaust steam from 
auxiliaries mixes with and heats the condensate from the main units. 



Desuperheater Spray Water 14,000 lb, 



10,000 lb. per Hour 



I 




Duplex Pump 



Fig. 358. 



HotWell Pump BoUer Feed Pump 

Make-up Water Evaporator System, Buffalo General Electric Co. 



The make-up water is boiled in an evaporator heated by high pressure 
steam and the resulting vapor passes directly to the barometric condenser 
in which it mixes with the auxiliary exhaust and thus becomes part 
of the feed water. For a full description of this interesting installation, 
consult ''The Connors Creek Plant of the Detroit Edison Company," 
C. F. Hirshfeld, Trans. A.S.M.E., Vol. 37, 1915. 

Fig. 358 gives a diagrammatic arrangement of the make-up water 
evaporator system of the River Station of the Buffalo General Electric 
Company, Black Rock, Buffalo, which is representative of the latest 
practice. Raw water is taken from the circulating water outlet of 
the main unit condensers and follows the course of the arrow heads from 
the open heater at left of the diagram, through the various apphances, 
to the economizer and thence to the boiler. Most of the impurities 
are precipitated in the evaporators from which they are discharged to 
waste. For complete details consult Power, Feb. 13, 1917, p. 202. 

285. Fuel Economizer. — Although any device which effects a saving 
in fuel is a fuel economizer the term ''fuel economizer" without quali- 
fication refers to a closed heater which receives its heat supply from 



608 



STEAM POWER PLANT ENGINEERING 



the flue gases. Two types of economizers are found in practice, (1) 
those which are independent of the boiler and (2) those which are in- 
tegral with the boiler and form a part of the heating surface. The 
independent type is the more common and is usually constructed of 
cast iron to obviate danger of corrosion. The integral type is usually 
constructed of wrought iron or steel tubes and is to all intents and pur- 
pose a part of the boiler proper. The present tendency toward higher 
boiler pressures makes the use of an economizer almost a necessity 
because of the otherwise high temperatures of the escaping flue gases; 
in fact, practically all modern large central stations are equipped with 
economizers. 

Fig. 359 gives a general view of a Green economizer, illustrating a 
typical flue gas heater. It consists of a series of cast-iron tubes 9 to 




Fig. 359. Green Economizer. 

10 feet in length and 4f inches in diameter, which are arranged ver- 
tically in sections of various widths across the main flue between boiler 
and chimney. When in position the sections are connected by top and 
bottom headers, and the headers are connected to branch pipes running 
lengthwise, one at the top and the other at the bottom. Both of the 
branch pipes are outside the brickwork which incloses the apparatus. 
The waste gases are led to the economizer by the ordinary flue from the 
boiler to the chimney, but a by-pass must be provided for use when the 
economizer is out of service for cleaning or for repairs. The feed water 
is forced into the economizer through the lower branch pipe nearest the 
point of exit of gases, and emerges through the upper branch pipe near- 
est the point where the gases enter. Each tube is encircled with a 
set of triple overlapping scrapers which travel continuously up and down 
the tubes at a slow rate of speed, the object being to keep the external 



FEED WATER PURIFIERS AND HEATERS 



609 



surfaces free from soot. The mechanism for working the scrapers is 
placed on top of the economizer, outside the chamber, and the motive 
power is suppUed either by a belt from some convenient shaft or small 
independent engine or motor. The power for operating the gearing 
varies from 1 to | horsepower per 1000 square feet of economizer sur- 
face, depending upon the number and length of tubes. The apparatus 
is fitted with blow-off and safety valves, and a space is provided at the 
bottom of the chamber for the collection of soot. For continuous plant 
operation the soot is automatically cleaned as shown in the illustration. 
This type of economizer is also used as an air heater for drying and 
heating purposes. The air heater is similar in design to the water 
heater with the exception of the direction of flow and size of tubes. 
The tubes in the air economizer are 3J inches internal diameter by 9 feet 



6*Saturated Headef 




Fi(i. 3G0. Typical Economizer Installation. 

in length, as against 4f inches internal diameter for the water econo- 
mizer. In the latter the water enters at the bottom header and passes 
out from the top header; in the former the air is forced by a fan first 
through one set of tubes and up through another set, and then down 
again, and so on until it leaves the heater. 

Fig. 361 shows a section through a 25,000-sq. ft. Badenhausen l^oiler 
as installed in the Highland Park plant of the Ford Motor Company 
and illustrates an economizer element integral with the boiler. Feed 
water enters drum 6, flows down the rear bank and enters the 
forward bank of tubes connecting drums 5 and 6. The economizer 
element is baffled so that the gases are forced to travel down the front 
bank and up the rear bank of tubes. The resulting difference in tem- 
perature creates a positive circulation of the water in the economizer 
element. The integral type of economizer is not commonly used in 



610 



STEAM POWER PLANT ENGINEERING 



this country but a modification of this arrangement, which appears 
to be the tendency in large central stations, is the subdivision of the 




Draft, Inches of Water 



Fig. 361. 25,000-sq. ft. Badenhausen Boiler with Economizer Element Integral 

with Heating Surface. 

heating surface so that each boiler has its own economizer. With large 
units the heating surface is often arranged in three or four sections. 

To maintain a constant velocity of the 
gases through the economizer passages each 
section is made narrower. 

Economizers have been installed in con- 
nection with chimney draft but the extra 
height of stack necessary to compensate 
for the reduction in draft caused by the 
lower temperatures of the gases and by 
the resistance of the tubes usually offsets 
the gain. In order to obtain an overall 
economy it is necessary to have some form 
of mechanical draft to force the gases 
through the economizer at proper speed. 
The loss in draft due to the reduction in 
temperature of the flue gases may be calculated as shown in para- 
graph 127. The loss in draft due to the resistance of the tubes varies 





















Uptake 


















\ 


















v 
















Entrance 


\ 


\, 














Ist Pass 




\ 


















\ 












2nd Pass 








\ 


















V 


s 






Exit 












N 


:^ 





Fig. 362. Pressure Drop through 
8500-sq. ft., 3-section Econo- 
mizer — Fan Draft. 



FEED WATER PURIFIERS AND HEATERS 611 

directly with the length of the economizer and as the square of the 
velocity. The pressure drop through an economizer 40 sections long 
with mean gas velocity of 1500 ft. per min. is approximately 0.25 in. 
water. This loss naturally varies with the design of the economizer. 
See curves in Fig. 362 for a specific example. 

286. Temperature Rise in Economizers. — The heat transfer in an 
economizer follows the same basic law as the heat transmission through 
any heating surface, viz. : 

SUd = w,ci (t - to), (245) 

= W2C2 {k - h), (246) 

in which 

S = total heating surface, sq. ft., 

U = mean coefficient of heat transmission, B.t.u. per hr. per 

sq. ft. per deg. mean temperature difference, 
d = mean temperature difference between the two fluids, deg. 
fahr., 
1^1 and W2 = weights, respectively, of the fluid to be heated and the 
flue gas, 
Ci and C2 = mean specific heats respectively of the fluid to be heated 
and the flue gas, 
^0 and t = initial and final temperature of the fluids to be heated, 

deg. fahr., 
<2 and ti = initial and final temperature of the flue gas, deg. fahr. 

By an analysis similar to that developed in paragraph 242 it may be 
shown that for either parallel or counter flow 

d = ^-i^ , (247) 

in which " tf 

ti,tf = initial and final temperature difference between the two 
fluids. 
By combining equations (245) to (247) and reducing (see Sibley 
Journal, Jan., 1916, p. 129) we have as an expression for the temperature 
rise in the feed water /„ _ /„ 

^ = ^ - 1 " ' (248) 

• u- V. 10" - 1 ^ 

m which 

X = temperature rise in the feed water, deg. fahr., 



n = 



W2C2 

SU (N - 1) 



2.3 wi 
Other notations as previously designated. 



612 



STEAM POWER PLANT ENGINEERING 



TABLE 108. 

AVERAGE MEAN COEFFICIENT OF HEAT TRANSFER IN ECONOMIZERS. 

(Clean Cast-iron Tube.s.) 

B.t.u. Per Sq. Ft. Per Deg. Fahr, Difference in Temperature, 





Mean Temperature Difference Between Flue Gas and Feed Water, Deg. Fahr. 


Velocity of the 




Gases, Ft. per Min. 














250 


275 


300 


350 


400 


500 


2 2 


2.3 


2.5 


2.7 


2.8 


1000 


3 


32 I 3.3 


3.4 


3.6 


1500 


3.6 


38 1 4.0 


4.2 


4.5 


2000 


4 


4.3 4.5 


4.7 


5.0 



Equation (248) applied strictly to counterflow which is the usual 
economizer practice. See Fig. (364). 

Example 56. Calculate the final feed- water and flue-gas tempera- 
ture for an economizer installation operating under the following con- 
ditions. Boiler heating surface 12,000 
sq. ft.; economizer surface 7500 sq. 
ft. ; initial feed-water temperature 100 
deg. fahr. and initial flue-gas tempera- 
ture 650 deg. fahr. when the boiler 
is operating at 100 per cent above 
standard rating; coal used, Illinois 
screenings, 11,400 B.t.u. per lb. 

























:g600 






































/60 





1 500 
§400 
















/ 



















^y 
















%»5> 


-< 















J 
















°.» 


IVfi 
































.^C' 




^2 


71 


2 200 
1 

e 100 






w 






-M 

^ 














^ 


jsS 




"•io 














^i 


5 














98 











































0123 4 56789 10 
Suriace, Thousands of Square Feet 

Fig. 363. Temperature of Flue 
Gas and Feed Water in an 
8000-sq. ft. Economizer — Fan 
Draft. 



-, - 


TIN 


- 


— 


— 


■ — 


— 


■- 


-" 


•~ 


— 


— 


-- 


— 


n 




, 1 


\ 
























-i 




i j 




s 


s 


fi 


, 
















i 
1 




" L 






V, 


N 


"o 


P^ 














1 


■', 


N 


'V, 






N 


\ 


"^ 














~i 


1 

1 




^ 


^ 




r 


^ 




\ 










1 
















pd 


H' 


., 






*v. 




—\ 




^ 1 










•^ 




P-> 












^ — r^ 




1 
















■~~ 






— 1 




- ^ v4. 


y. 


:■ l_. 






















.. 




I'o l^ 



Fig. 364. Counter Current Flow. 



It has been shown (paragraph 21) that the theoretical weight of air 
per lb. of any coal is approximately 7.5 lb. per 10,000 B.t.u. Therefore 
for the coal specified, 

1.14 X 7.5 = 8.65 lb. = theoretical air requirements per lb. of coal. 

Assuming an air excess of 50 per cent at maximum load and allowing 
15 per cent for ash the probably actual weight of flue gas per lb. of coal 
= 1.5 X 8.65 + 0.85 = 13.8 lb., or in round numbers 14 lb. 

Since the evaporation at rating is equivalent to 3.45 lb. from and at 
212 deg. per sq. ft. heating surface per hr., at 100 per cent overload the 
total weight of water, w, fed to the boiler is 

w = 2X 12,000 X 3.45 = 82,800 lb. per hr. 



FEED WATER PURIFIERS AND HEATERS 



613 



Assuming an overall efficiency of 75 per cent the weight of coal re- 
quired is 

970.4x82,800 ^.^^i, , 

11,400X0.75 -Q^OOlb.perhr. 

The total weight of flue gas, W2, is 

W2 = 9400 X 13 - 131,600 lb. per hr. 

Assume the mean specific heat of the water to be unity and that of 
the flue gas to be 0.25. 

Assume U = 4.25, which is an average value for a modern economizer 
with initial flue gas temperature of 650 deg. fahr. Substituting these 
values in equation (248), 

TABLE 109. 

ECONOMIZER PROPORTIONS IN MODERN CENTRAL STATIONS. 



Name of Plant. 



Buffalo General Electric 

*Cleveland Municipal Plant, 53rd Street 
Station 

Commonwealth Edison Co., Fisk Street 

Station 

Northwest No. 3 

fDelray, No. 1 

Public Service, Joliet, 111 

Public Service, New Jersey, Essex Sta- 
tion 

Regina, Sask., Can 



Size of 
Boiler Unit 
Nominal 
Horse- 
power. 



1140 

1013 

1225 

1220 

483 

992 

1373 
500 



Boiler 

Heating 

Surface Per 

L'nit, 

Sq. Ft. 



11,400 

10,134 

12,250 

12,200 

4,830 

9,919 

13,730 
5,000 



Economizer 

Surface Per 

Boiler Unit, 

Sq. Ft. 



9435 

5400 

8500 
6566 
4896 
6730 

7750 
2320 



Ratio 

Economizer 

to Boiler 

Surface. 



0.825 

0.525 

0.692 
0.540 
0.490 
0.679 



0.564 
0.464 



• One economizer for 5 boilers. 



t One economizer for 2 boilers. 



TV = ^^ = 



82,800 X 1 



= 2.52. 



W2C2 131,600X0.25 

SU (N - 1) 7500 X 4.25 (2.52 - 1) 






2.3 X 82,800 
650 - 100 



= 0.254. 



N -I 



-\-N 



2.52 - 1 



= 126 deg. fahr. 



+ 2.52 



lO'*- 1 ' ^' 100 254- 1 

Since x = t — U, the final temperature of the feed water is 
^ = 126 + 100 = 226 deg. fahr. 

The heat absorbed by the feed water must be equal to that given up 
by the flue gas, or 

WiCi {t - to) = W2C2 {t2 - t), (249) 

from which 

(^' = "^' = iV. (250) 

t — to W2C2 



614 



STEAM POWER PLANT ENGINEERING 



Substituting the known quantities in equation (250) 

650 - ti 



226 - 100 



= 2.52, 



or 



ti = 337.0 deg. fahr. = final temperature of the flue gas. 
For parallel flow as in Fig. 365 the final flue gas temperature may be 



2 


"H 


^ 


~' 


— 


— 


~' 


"■ 


■" 








— 


' 





r-f 




1 




■^ 


^. 




pi, 1 












1 


1 














^. 












1 














""■ 


~:^ 


■" 










i 


•1 


























■ tf 


ir 


1 










aV 


gaV 















r 






1 








r 




— ■ 














1 ' 




■ \< 


^ 
























1 


1 


J 


























1 




^\ 


_. 








-.. 
















1 s 





Fig. 365. Parallel Current Flow. 



calculated from the following formula which has deauced from equations 
(245) to (247). 

a 



k = 



b a 

i^ + b 



(251) 



in which 



a = — U + ^0, 
W2C2 

6 = ^ + 1. 

W2C2 



m 



W2C2 



Other notations as previously designated. 

387. Value of Economizers. — The general conclusions drawn from 
current practice is that an economizer installation results in: 

(1) A saving in fuel ranging from 7 to 20 per cent. 

(2) A very small gain and often an actual loss in overall economy 
when installed in connection with feeble chimney draft and underloaded 
boilers. 

(3) A substantial overall gain in economy where the boilers are 
forced and mechanical draft is employed. 

(4) Maximum overall economy when the boilers are forced far 
above their rating and the auxiliaries are electrically driven and pure 
feed water is available. 

(5) Decreased wear and tear on the boilers due to the high feed-water 
temperature. 

(6) A large storage of hot water for sudden peak demands. 



FEED WATER PURIFIERS AND HEATERS 



015 



TABLE 110. 

ECONOMIZER PERFORMANCES. 



Number of 
Plant. 



Number of 
Economizer 

Tubes 
Installed. 



Temperatures, Deg. Fah. 



Gases 

Entering 

Economizer 



Gases 

Leaving 

Economizer 



Fluid 

Entering 

Economizer 



Fluid 

Leaving 

Economizer 



Rise in 

Temperature 
of Fluid. 



Actual 
Saving in 

Fuel, 
Per Cent. 









Water Heater. 








1 


160 


435 


279 


84.2 


196.2 


112.0 


12.5 


2 




416 
620 


254 
293 


40.0 
101.0 


185.4 
237.0 


125.4 
136.0 


13.8 


3 


960 


18.3 


4 


520 


548 


295 


96.0 


200.0 


104.0 


9.2 


5 


520 


603 


325 


93.5 


203.8 


110.3 


9.7 


6 


384 


368 


245 


103.0 


202.6 


99.6 


12.4 


7 


448 


537 


326 


71.2 


203.4 


132.2 


17.5 



Air Heater. 



72 


301 


257 


70.0 


152.0 


240 


512 


319 


54.0 


201.6 


96 


557 


376 


41.0 


200.0 


192 


417 


369 


74.0 


210.0 



82 
147.6 
159.0 
136.0 



9.0 
14.0 



Compiled from *' The Book of the Economizer," 1912, published by the Green Engineering Co. 

288. Factors Determining Installation of Economizers. — Some of the 
more important factors to be considered before installing an economizer 
are: 

(1) Temperature of the flue gas. The higher the temperature of 
the flue gas the greater will be the thermal saving. With the standard 
type of boiler operating with high pressures, 300 lb. per sq. in. or more, 
economizers are practically indispensable. See Table 36 for flue gas 
temperatures incident to boiler overloads. 

(2) Initial temperature of the feed water. With electrically driven 
auxiliaries exhaust steam is not available for heating the feed water and 
an economizer is desirable. Even with initial temperature as high as 
200 deg. fahr. overall economy may result from the use of an economizer. 

(3) Purity of the feed water. With impure feed water the formation 
of scale within the tubes may seriously affect the efficiency of heat 
transmission and the cost of cleaning may prove excessive. Internal 
corrosion may also be caused by impure feed water. 

(4) Minimum temperature of the flue gas. The flue gas tempera- 
ture should not be lowered below the dew point since the conden- 
sation of the vapor content may cause the soot to adhere to the tubes 
and render its removal a costly problem. An average minimum is 
240 deg. fahr. With coals high in sulphur content the moisture forms 
sulphuric acid which corrodes the tubes. 



P 



616 STEAM POWER PLANT ENGINEERING 

(5) Increased capacity due to the additional heating surface. 

(6) Cost of additional building space. With the independent type 
of economizer this is of secondary importance. 

(7) Cost of producing the draft. For chimney draft this means cost 
of the extra height of stack necessary to overcome the loss in draft. 
This may range from 20 to 40 per cent of the total cost of the chimney. 
In the modern mechanical draft installation the power required to 
operate the fan ranges from one per cent to four per cent of the main 
generator output. 

(8) First cost. Economizers cost approximately $1.25 per sq. ft. 
of surface for pressures under 250 lb. per sq. in., though the cost natu- 
rally varies with the cost of raw material. 

(9) Boiler pressure. Cast-iron superheaters are used for working 
pressures as high as 400 lb. per sq. in. but the cost increases rapidly with 
increase in pressure above 250 lb. per sq. in. It is doubtful if cast iron 
will be used in projected new plants where pressures of 500 lb. or more 
are being seriously considered. 

289. Choice of Feed-water Heating System. — The heating of feed 
water and Its dehvery to the boiler in the most economical manner 
is a problem involving such a large number of combinations that a 
general analysis is impracticable. The following discussion of a spe- 
cific case will give some idea of the manner in which this problem may 
be attacked. 

Example 57. Determine the most economical manner of heating the 
feed water for a power plant of 1000 horsepower operating under the 
following conditions: Schedule 10 hours per day and 310 days per year; 
load factor on the ten-hour basis 0.8; cost of coal $2.50 per ton of 2000 
pounds; heat value of the coal 13,500 B.t.u. per pound; average boiler 
efficiency 65 per cent; engines use 20 pounds of steam per i.hp-hour; 
steam pressure 150 pounds absolute; temperature of cold water 60 
degrees; vacuum 26 inches referred to 30-inch barometer; interest 
5 per cent; depreciation 8| per cent; maintenance 1 per cent; in- 
surance J per cent; taxes 1 per cent; total charges 16 per cent; charges 
for attendance and maintenance assumed to be the same in each case 
and credit for the chimney assumed to offset debit for economizer 
space. Many of the influencing conditions are left out for the sake of 
simplicity. 

The most likely combinations are 

(1) Atmospheric, all auxiliaries steam driven, water taken from cold 

well. 

(2) Same as (1) except that water is taken from hot well. 

(3) Economizers, auxiliaries electrically driven, chimney draft, water 

from cold well. 

(4) Vacuum heater, economizer, and electrically driven auxiliaries, 

fan draft. 



FEED WATER PURIFIERS AND HEATERS 617 

(5) Vacuum heater, atmospheric heater, and steam auxiharies. 

(6) Atmospheric heater, economizer, steam auxiharies, fan draft. 

(7) Vacuum and atmospheric heaters, economizers, steam auxiharies, 

and electrical fan. 

(8) Vacuum, atmospheric heater, economizer, and chimney draft, 

auxiliaries operating condensing except feed pumps and stoker 
engines which exhaust into the atmospheric heater. 

The difference between the total heat furnished by the boiler and 
the heat returned in the feed water is the net heat put into the steam 
by the boiler. Evidently the system which shows the least net heat 
required to produce one horsepower will be the most economical as 
far as coal consumption is concerned, although not necessarily the 
cheapest when both operating and fixed charges are considered. 

Prices vary so much that it is practically impossible to give costs of 
installations which will bear criticism and the prices taken in this prob- 
lem are approximate only. 

Case I. 

Atmospheric heater, auxiliaries steam driven, feed from cold well. 
This arrangement and that of Case II are the most common in power 
plants of this size. 

The power consumption of the auxiliaries operating non-condensing 
varies from 8 to 12 per cent of the total power developed. Assume 
it to be 10 per cent. 

The temperature of the feed water leaving the heater may be deter- 
mined by equation (237). 

^ ^0 + 0.9 .S (X + 32) 
1+0.9.S 
Substituting S = 0.10, X = 1146, ^o = 60, 

^ 60 + 0.9 X 0.10(1146 +32) 

1 + 0.9 X 0.10 
= 152. 

The net heat furnished by the boiler to produce one indicated horse- 
power-hour in the engine is evidently the heat necessary to raise 20 + 10 
per cent of 20 = 22 pounds of water from 152 deg. fahr. to steam at 
150 pounds pressure; i.e., the net heat furnished is 

22 X 1071.2 = 23,564 B.t.u. 

Now, 1 i.hp. = 2546 B.t.u. 

Therefore the heat efficiency of this arrangement is 

2546 ^„^ 
23^ =10.8 per cent. 

Probable First Cost. 

Steam pumps $400 . 00 

Condenser with steam-driven air and circulating pumps 3000.00 

1000-horsepower open heater 480 . 00 

Piping 1200.00 

$5080.00 



618 STEAM POWER PLANT ENGINEERING 



Fuel Consumption. 

Average horsepower-hours per year = 1000 (rated horsepower) X 0.8 (curve 
load factor) X 310 (days per year) X 10 (hours per day) = 2,480,000. 

Pounds of coal per i.hp-hour = net heat furnished per i.hp-hour -J- net heat ab- 
sorbed by the boiler per pound of coal = 23,564 -=- (13,500 X 0.65) = 2.68. 

^ 2,480,000 X 2.68 „„„ 
Tons per year = -^ ^qqo ^ ^' 

Fuel and Fixed Charges. 

Fuel, 3323 tons at $2.50 $8308.00 

Fixed charges, 16 per cent of $5080 812.00 

$9120.00 

Case II. 

Same as Case I, except that feed is taken from the hot well. This 
arrangement is possible only when the condensing water is suitable 
for feed purposes. 

Assume the temperature of the water from the hot well as it enters 
the heater to be 110 degrees. 

The temperature of the feed water leaving the heater will then be 
198 degrees (from equation (237)). 

Net heat furnished = 22 X 1025.2 = 22,554 B.t.u. 

Efficiency = ^^ . = 11.3 per cent. 

22 554 
Pounds of coal per i.hp-hr. = ^^^^^^^^^^ = 2.62. 

^ 2,480,000 X 2.62 ^_ .» 
Tons per year = -^ ^^qoo ^ ^ ^' 

Fuel and Fixed Charges. 

Fuel, 3248 tons at S2.50 $8120.00 

Fixed charges (same as Case I) 812.00 

$8932.00 

Case III. 

Economizers, auxiliaries electrically driven, chimney draft, water 
from the cold well. 

Practice gives an average of 3 per cent of the main engine output as 
the power required to operate the electrical auxiliaries in a plant of this 
size. 

The temperature rise of the feed water leaving the economizer is 
found to be 119 deg. fahr. (equation 248). 

Temperature of feed water entering boiler = 119 + 60 = 179 degrees. 

Net heat furnished = (20 + 3 per cent of 20) X 1044.2 = 21,510 

B.t.u. 

2545 
Efficiency = =11.8 per cent. 

^l,olU 



I 



FEED WATER PURIFIERS AND HEATERS 619 

Probable First Cost. 

Economizers S3500.00 

Motor feed pump 600 . 00 

Condenser with electrically driven air and circulating pump . . . 6000 . 00 

Piping and wiring 1000 . 00 

$11,100.00 

Fuel Consum-plion. 

Pounds of coal per i.hp-hr. = .o :^c\c^ ^, n a- = 2.45. 

13,500 X 0.6o 

^ 2,480,000 X 2.45 „„„o 
Tons per year = ^ ^qoo " ^^^^' 

Fuel and Fixed Charges. 

Fuel, 3038 tons at $2.50 $7595.00 

Fixed charges, 16 per cent on $11,100 1776 00 

$9371.00 

Case IV. 

Vacuum heater, economizer, electrically driven auxiliaries, fan draft. 

The vacuum heater may be relied upon to raise the temperature of 
the feed water to 110 degrees. 

The economizer will increase this 107 degrees (from equation (248)), 
giving the feed water a temperature of 217 degrees as it enters the 
boiler. 

The electrical fan for the mechanical-draft system will require ap- 
proximately 2 per cent of the main system engine power, making a 
total of 3 + 2 = 5 per cent for all auxiliaries. 

Net heat furnished = (20 + 5 per cent of 20) X 1006.2 
= 21,130 B.t.u. 

2545 
Efficiency = = 12.05 per cent. 

Probable First Cost. 

For the sake of simplicity it is assumed that the high first cost of the 
chimney plus its low depreciation and maintenance will offset the low 
first cost of the mechanical-draft system plus its higher maintenance 
and depreciation charges: 

Economizers $3500 . 00 

Motor feed pump 600.00 

Motor-driven pumps and condenser 6000 . 00 

Motor-driven fan 750 . 00 

Piping and wiring 1200 . 00 

Vacuum heater 200.00 

$12,250.00 
Fuel Consumption, 

21 130 
Pounds of coal per i.hp-hr. = Y 3, 500 X 0.65 = ^^^^' 

^ 2,480,00 X 2.41 ____ 
Tons per year = -^ ^^qoo " ^^^' 



620 



STEAM. POWER PLANT ENGINEERING 



Fuel and Fixed Charges. 

Fuel, 2988 tons at $2.50 $7470.00 

Fixed charges, 16 per cent of $12,250 1960.00 

$9430.00 

In like manner Cases V, VI, VII, and VIII have been treated and are 
tabulated in the summaries. 

SUMMARY (1). 



Case. 


Temperature 
of Feed 
Water. 


Power ' 
Consumed by 
Auxiliaries. 


Efficiency. 


First 
Cost. 


Fuel Cost 
per Year. 


Cost of 
Operation 
per Year. 


I 


Degrees F. 
152 
198 
179 
217 
208 
294 
290 
270 


Per Cent, 
10 
10 

3 

5 

10 
14 
10 

8 


Per Cent. 
10.8 
11.3 
11.8 
12.05 
11.4 
12 

12.2 
12.3 


$5,080 
5,080 
11,100 
12,250 
5,280 
9,000 
9,300 
8,250 


$8,308 
8,120 
7,595 
7,470 
7,900 
7,750 
7,380 
7.075 


$9,120 


II 


8,932 


Ill 


9,371 


IV 


9,430 


V 


8,744 


VI 


9,190 


VII 

VIII 


9,570 
8,395 



SUMMARY (2). 



Case. 


Efficiency. 


First Cost. 


Fuel. 


Cost per Year. 


I 


8 
7 
6 
3 
5 
4 
2 
1 


1 
1 
6 
7 
2 
4 
5 
3 


8 
7 
4 
3 
6 
5 
2 
1 


4 


II 


2 


Ill 


6 


IV 


7 


V 


3 


VI 


5 


VII 


8 


VIII 


1 







Summary (2) gives the ranking ; thus : Case I is eighth in point of effi- 
ciency first in cheapness of installation; eighth in yearly cost of fuel; 
and fourth in yearly cost of operation. Case VIII is apparently the 
best arrangement for the given conditions. 

Bleeding Turbines to Heal Feed Water: Power, May 15, 1917, p. 652. 



PROBLEMS. 

1. Determine the amount of soda ash and lime necessary to soften 10,000 gallons 
of water as per analysis. Col. 2, Table 98. 

2. In a certain plant it costs 30 cents per 1000 lb. to evaporate water from feed 
temperature of 60 degrees to steam at 115 lb. abs. and 50 deg. superheat; required 
the saving in per cent if the feed water is heated by exhaust steam to 210 deg. fahr. 

3. A 2000-kw. turbo-generator plant uses 18 lb. steam per kw-hr., initial pressure 
140 lb. abs., back pressure 3 in. abs., superheat 100 deg. fahr., temperature of the 



FEED WATER PURIFIERS AND HEATERS 621 

condensate 100 deg. fahr.; auxiliaries develop 100 hp. and use 30 lb. steam per hp- 
hr. (non-condensing), initial pressure 115 lb. abs., steam dry at admission; re- 
quired the temperature of the feed water if the auxiliary exhaust is discharged into 
an open heater, 

4. Required the tube surface necessary for a closed heater suitable for the con- 
ditions in Problem 3. Assume U = 350. 

5. If the tubes are ^ inch inside diameter, required the total length of water 
travel for the conditions in Problem 4, assuming a water velocity through the tubes 
of 120 ft. per min. 

6. Calculate the final feed-water and flue-gas temperatures for an economizer 
installation operating under the following conditions: Boiler heating surface 10,000 
sq. ft., economizer surface 6500 sq. ft., initial feed-water temperature 120 deg. fahr., 
initial flue-gas temperature 700 deg. fahr. when the boiler is operating at 150 above 
standard rating; coal used, Illinois washed nut, 13,500 B.t.u. per lb. 



CHAPTER XIII 



PUMPS 



390. Classiflcation. — Pumps used in connection with steam power 
plants may be conveniently classified under five groups according to 
the principles of action. 

1. Piston pumps, in which motion and pressure are imparted to 
the fluid by a reciprocating piston, plunger, or bucket. The action is 
positive and a certain definite amount of fluid is handled per stroke 
under predetermined conditions of pressure and velocity. 
I 2. Centrifugal pumps, in which the fluid is given initial velocity and 
pressure by a rotating impeller. The action is not positive, as the 
amount of fluid discharged is not necessarily proportional to the im- 
peller displacement. 

3. Positive displacement rotary pumps, in which motion and pressure 
are imparted to the fluid by a rotating impeller or screw. The volume 
discharged is practically equal to the impeller displacement regardless 
of pressure. 

4. Jet pumps, in which velocity and pressure are imparted to the 
fluid by the momentum of a jet of similar or other fluid. The ordinary 
steam injector is the best known of this group. 

5. Direct-pressure pumps, in which the pressure of one fluid acts 
directly on the surface of another fluid, thereby imparting all or part 
of its energy to the latter. The pulsometer is an example of this type. 

These groups may be variously subdivided as follows: 



Piston 



Direct-acting 



( Simplex. 
I Duplex. 
( Simplex . 



Fly-wheel < Duplex , 

Power driven . . ( Triplex , 

c-'"f'«''i IKrt^^„e:::::::lS-ri:; 

Power driven . . \ Forcing . . . . 

/ Lifting 

\ Injector { Positive . . . . 

] Ejector } Automatic. . 

j Pulsometer Lifting 

I Air-lift Lifting 



Rotary 

Jet 

Direct pressure . 



Air. 

Vacuum. 
Forcing. 
Lifting. 



Vacuum. 
Forcing. 
Lifting. 



Piston or plunger pumps are the most common in use. Small boiler- 
feed pumps, city waterworks pumps and force pumps are ordinarily 
of this type. In the direct-acting type, Fig. 367, the water plunger 

622 



PUMPS 623 

and steam piston are secured to a single piston rod and the steam pres- 
sure is transmitted directly to the water. There is no flywheel, connect- 
ing rod, or crank. The velocity of the delivery is proportional to the 
resistance offered by the water; when the resistance equals the forward 
effort of the steam pressure the pump stops. This class of pump is 
well adapted for boiler-feeding purposes, since it may be operated as 
slowly as suits the requirements of feeding by simply throttling the 
discharge. The steam consumption is very large in proportion to the 
work performed, since the steam is not used expansively. 

Flywheel pumps, Figs. 380, 428, are ordinarily classified as pumpmg 
engines. In this class steam may be used expansively, as sufficient 
energy is stored in a flywheel to permit the drop in steam pressure during 
expansion. These pumps find wide application in city waterworks, 
elevator plants, and the like, where high duty is required. They 
are little used as stationary boiler feeders, but are used to some extent 
in river-boat practice and in plants operating continuously for long 
periods at comparatively steady loads. Practically all sizes of dry-air 
pumps and a number of large jet condenser pumps are of this type. 

Piston pumps, Fig. 387, driven by gearing or belting are ordinarily 
classified as power-driven pumps. The driving power may be steam 
engine, electric motor, or gas engine. The single-cylinder machine is 
often designated as a '^ simplex" power-driven pump, the two-cyUnder 
as a "duplex," the three-cylinder as a "triplex," and so on. 

Centrifugal pumps, Fig. 415, are supplanting to a considerable extent 
the present type of piston pump for many uses. Though particularly 
adapted for low heads and large volumes, they are used in many situ- 
ations requiring extremely high heads. They are not as efficient as 
high-grade pumping engines, but the extremely low first cost fre- 
quently offsets this disadvantage, and they are much used in connection 
with dry docks, irrigating plants, sewage systems, and as circulating 
and vacuum pumps in condensing plants. 

Rotary pumps, Fig. 424, are employed to a limited extent in the 
same field as the centrifugal pump. Being positive in action, they 
permit of a much lower rotative speed for the same dehvery pressure. 

Jet pumps. Fig. 391, are seldom used as pumps in the ordinary sense 
of the word, on account of their extremely low efficiency, but are fre- 
quently employed for discharging water from sumps. Their greatest 
field of application lies in boiler feeding and in this respect their effi- 
ciency is comparable with that of the average piston pump. A re- 
cently developed multi-jet air pump gives great promise of superseding 
the present type of dry-air pump for vacuum purpose. See para- 
graph 309. 



624 



STEAM POWER PLANT ENGINEERING 



Direct-pressure pumps operated by steam, such as the '^pulsometer, " 
Fig. 430a, are used principally for pumping out sumps, surface drains, 
and the like, where the operation is intermittent. Direct-pressure 
pumps of the air-lift type. Fig. 431, are quite common and are used a 
great deal in situations where water is to be pumped from a number of 
scattered wells. 

291. Boiler-feed Pumps, Direct-acting Duplex. — Figs. 366 and 367 
illustrate a typical duplex boiler-feed pump, which consists virtually of 
two direct-acting pumps mounted side by side, the water ends and the 



Air 
Chamber 



Dischargo 




Suction 



Fig. 366. Typical Duplex Pump. 



steam ends working in parallel between inlet and exhaust pipe. The 
piston rod of one pump operates the steam valve of the other through 
the medium of bell cranks and rocker arms. The pistons move alter- 
nately, and one or the other is always in motion, the flow of water being 
practically continuous. 

In general construction the steam pistons and valves are similar to 
those of steam engines. The valves in duplex pumps, however, have 
no lap. In order to reduce the valve travel to a minimum, and 
still have sufficient bearing surface between the steam ports and the 
main exhaust ports to prevent the leakage of steam from one to the 
other, separate exhaust ports are provided which enter the cyhnder 
at nearly the same point as the steam ports. This arrangement offers 



PUMPS 



625 



a simple means of cushioning the piston by exhaust steam, thus pre- 
venting it from striking the cy Under heads at the ends of the stroke. 
The valves of the duplex pump having no lap would, if connected 
rigidly to the valve stem, open one port as soon as the other had been 



DISCHARGE 




Fig. 367. Section through a Typical Duplex Boiler-feed Pump. 

closed, at about mid-stroke of the piston, thus cutting down the stroke 
to about one fourth the usual length. To obviate this difficulty the 
valves are given considerable lost motion by allowing sufficient clear- 
ance between the lock nuts on the valve stem; the latter, therefore, 
imparts no motion to the valve until the piston operating it has nearly 
completed the stroke. The lost motion between valves and lock nuts 
renders it impossible to stop the pump in any position from which it 
cannot be started by simply admitting steam, and therefore the pump 
has no dead centers. When one piston moves to the end of the stroke 
it pulls or pushes the opposite valve to the end of its travel; then when 
the piston starts back to the other end of its stroke the valve remains 
stationary, owing to the lost motion, until the piston has completed 
about one half the stroke. During this time the opposite piston has 
completed a full stroke and the valve operated by it will have opened 
the steam port wide, so that while one valve covers both steam ports 
the other is at the end of its travel. In some makes of pumps the stem 
is rigidly attached to the valves, the lost motion being adjusted outside 
the steam chest as shown in Figs. 368 and 369, which represent two 
common constructions of duplex valve gear. 

Fig. 370 shows the valve and piston in the position occupied at the 



626 



STEAM POWER PLANT ENGINEERING 



commencement of the stroke. At one end of the valve the steam port 
P is open wide and at the opposite end the exhaust port E is open wide. 



go|g J 




Valve Stem 



Piston Rod 



Fig. 368. 




^ Pfston Rod 




Fig. 370. 



When the piston nears the opposite end of the stroke and reaches the 
position shown in Fig. 371 the steam escape through the exhaust port 

^ ,,^, E is cut ofT by the piston, and 

^^,f--^|[i== i^ ^^_ _ ^^^ gince the steam port is closed, 

the remaining steam is com- 
pressed between the piston and 
cyhnder head, thus arresting 
the motion of the piston grad- 
ually without shock or jar. 

The construction of the water 
end of single-cylinder and du- 
plex pumps is practically the 
same; any sHght differences 
which may be found are con- 
fined to minor details which in no way affect the general design or 
operation of the pump. The 
piston is double acting, the 
single-acting cylinder being con- 
fined to power pumps or to 
steam pumps intended for very 
high pressures. In the old-style 
pumps it was the custom to use 
one large valve with a lift suf- 
ficient to give the required pas- 
sage, but in modern practice 
the required area is divided 
among several small valves, so ^^^- ^^^• 

that each one is easily and cheaply removed in case of accident or 
wear, and slip is lessened.* 

* The modern Riedler pump is an exception. See Engineer, U. S., Nov. 15, 1907, 
p. 1040. 



^g^^^ ^ 




PUMPS 



627 



The valves are carried by two plates or decks, the suction valves 
being attached to the lower plate and the delivery valves to the upper 
one, as shown in Fig. 368. 

The valves in practically all boiler-feed pumps are of the flat disk 
type, Fig. 372, held firmly to the seat 
by conical springs and guided by a bolt 
through the center. 

All pumps are provided with an air 
chamber on the discharge side, which 
acts as a cushion for the water, prevents 
excessive pounding, and insures a uni- 
form flow. Fig. 373 shows a section 
through the steam end of a compound 
duplex pump. 

292. Feed Pumps with Steam-actuated 
Valves. — Single-cyhnder direct-acting 
pumps. Fig. 374, are ordinarily operated 
by steam-actuated valves. The steam enters the chest C and passes to 
the left through the annular opening A formed between the reduced 
neck of the valve and the bore of the steam chest. It is thus projected 
against the inside surface of the valve head H before escaping through 
the port P and passing to the cyhnder. Both the pressure and impulse 




Fig. 372. A Typical Pump 
Disk Valve. 







Fig. 373. Section through Steam Cylinders of a Typical Compound Duplex Pump. 



due to velocity acting on the valve head H tend to close or restrict the 
admission port by forcing the valve to the left. On reaching the cyhnder 
and forcing the piston X toward the right, the pressure of the steam upon 
the opposite side of the valve head H is pressing the valve to the right, 
a movement which would give the admission more port opening at A 



628 



STEAM POWER PLANT ENGINEERING 



and deliver more steam to the cylinder. The valve then holds a po- 
sition depending upon the relative intensity of the two pressures, which 
tend to move it in opposite directions, the admission steam, tending to 
close the valve, and cyUnder steam, tending to open the valve wider. 




steam Supply 



Ail? 




Fig. 374. Marsh Boiler-feed Pump, A Typical Steam-actuated Valve Gear. 

The steam valve, therefore, is always in a balanced position. The 
steam piston is grooved at the center, forming a reservoir for live steam 
R which is supplied from the upper chamber of the steam chest by pas- 
sage E to the cylinder cap S, and thence by tube M and the hollow 
piston rod V. The steam in this annular piston space reverses the steam 
valve by pressing alternately against the outer surfaces of the valve 
heads H through the connecting passages 0, near each end of the cyl- 
inder. The tappets T are for the purpose of moving the valve by hand 
in case it fails to move automatically. Steam-actuated valves are not 
as positive in action as mechanically operated valves, and hence are 
little used in situations where positive action is essential, as in fire- 
pump service. 

293. Air and Vacuum Chambers. — Air chambers in piston pumps 
are for the purpose of causing a steady discharge of water and of re- 
ducing excessive pounding at high speeds by providing a cushion for 



PUMPS 



629 




r^ 



the water. The water discharged under pressure compresses the air 
in the air chamber somewhat above the normal pressure of discharge 
during each stroke of the water piston, and when the piston stops mo- 
mentarily at the end of the stroke the air expands to a certain extent 
and tends to produce a uniform rate of flow. 

The volume of the air chamber varies 
from 2 to 3J times the volume of the 
water piston displacement in single-cylinder 
pumps, and from 1 to 2J times in the du- 
plex type. High-speed pumps are provided 
with air 'chambers of from 5 to 6 times the 
piston displacement. The water level in 
the air chamber should be kept down to 
one fourth the height of the chamber. In 
slow-running pumps sufficient air may be 
carried into the pump chamber along with 
the water, but with high speeds a large 
part of the air will be discharged, and air 
must be forced into the chamber by mechanical means. The larger 
the chamber the more uniform will be the discharge pressure. 

Vacuum chambers are frequently provided for the purpose of main- 
taining a uniform flow of water in the suction pipe and assisting in the 
reduction of slip. Such chambers should be of slightly greater volume 
than the suction pipe and of considerable length rather than diameter. 



Fig. 



375. Forms of Vacuum 
Chambers. 




Different Arrangements of Vacuum Chambers. 



Fig. 375 illustrates two designs commonly used. The one in Fig. 375 (B) 
should be placed in such a position as to receive the impact of the 
column ©f water in the suction pipe as illustrated in Fig. 376 (A), (B) 
and (C). The chamber illustrated in Fig. 375 (A) should be placed in 
the suction pipe below but close to the pump. 



630 



STEAM POWER PLANT ENGINEERING 



294. Water Pistons and Plungers. — In cold-water pumps the water 
pistons are usually packed with some kind of soft packing. Fig. 377 (A) 
shows the details of a piston with square hydraulic packing. The body 
E is fastened to the piston rod by nut C; packing is placed at D, and 




Fig. 3'i 



(B) (C) 

Types of Water Pistons. 



follower F is forced up by the nut B and locked by nut A. For large 
sizes the design is the same except that the follower is set up by a num- 
ber of nuts near the edge. In hot-water pumps the pistons are often 
packed by means of metallic piston rings R, R, Fig. 377 (C), similar 
to those in steam pistons, or merely by water grooves G, G, Fig. 377 (B). 
The water end is often fitted with a plunger instead of a piston, as in 



^H 



-^^^ 




Fig. 378. Plunger with Metal Packing Ring. 

Figs. 378 to 380. The piston is more compact, but the plungers do 
not require a bored cylinder, so that the first cost is not materially 
different. 

Fig. 378 shows a plunger with metal packing ring. When leakage 
becomes excessive it is necessary to renew the ring, which is readily 
removed. 



PUMPS 



631 



In Fig. 379 the plunger is packed with hydrauhc packing as in the 
follower type of pump piston. The great difficulty with the above 
types of piston and plunger is in keeping the packing tight or in know- 




FiG. 379. Plunger with Hydraulic Packing. 

ing when it is leaking, and the trouble necessary to replace the packing. 
The outside packed plunger, Fig. 380, obviates these disadvantages to a 
great extent, since leakage is readily detected and repacking is performed 




Fig. 380. Horizontal Flywheel Pump with Outside Packed Plunger. 

without removing the cyhnder heads. In dirty, dusty locations, how- 
ever, the piston pump or inside packed plunger is to be preferred, since 
the abrasive action of the dust renders outside packing difficult. Fig. 
380 illustrates a high-duty elevator pump with outside packed plunger. 



632 



STEAM POWER PLANT ENGINEERING 



395. Performance of Piston Pumps. — Direct-acting pumps as a class 
are wasteful of fuel and low in efficiency, due largely to the non-ex- 
pansive use of steam. The average small duplex boiler-feed pump uses 
from 100 to 200 pounds of steam per i.hp-hr., depending upon the 
speed, and the mechanical efficiency varies from 50 per cent to 90 
per cent. When new and in proper working condition the mechanical 
efficiency is seldom less than 85 per cent; but such pumps, as a rule, 
are given scant attention, and the average efficiency is not far from 
65 per cent. The term ''mechanical efficiency" in this connection re- 
fers to the ratio of the actual water horsepower to the indicated horse- 
power of the steam cylinder. The loss includes the shp of the piston 



400 



a 200 

o 

I 



a 100 



















I 






Effect of Speed 

on the 

Economy of Small Direct- Acting 

Steam Pumps 




v 






A 16x 
B 12x 


10 X 12 

7Xxl2 


Duplex 
Simplex 






\ 


\b 




\ 




















1 A 


• 











50 75 100 125 150 

Number of Single Strokes Per Minute 

Fig. 381. 



175 



200 



and valves. A steam consumption of 150 pounds per i.hp-hour with 
mechanical efficiency oi 65 per cent is equivalent to a power consump- 
tion of about 5 per cent of the rated boiler capacity, although if the 
exhaust steam is used for feed-water heating the actual heat consump- 
tion may be but 1 to 1.5 per cent. Compound direct-acting pumps 
running non-condensing use from 50 to 100 pounds of steam per i.hp- 
hour. Single-cylinder flywheel pumps of the slow-speed type, running 
non-condensing, use about 50 pounds of steam per i.hp-hour. Multi- 
cylinder flywheel pumps of the high-duty type use about 25 pounds 
per i.hp-hour when running non-condensing, and as low as 10 pounds 
when operating condensing. High-grade direct-connected motor-driven 
power pumps have a mechanical efficiency from line to water load, at 
normal rating, of about 80 per cent. The efficiency of geared pumps at 
normal rating varies with the character of the gearing and the degree of 
speed reduction, and may range anywhere from 40 to 70 per cent. 



PUMPS 



633 



The steam consumption of all direct-acting boiler pumps decreases 
with the increase in speed. This is illustrated by curve J5, Fig. 381, 
plotted from the tests of a 12-in. by Tj-in. by 12-in. direct-acting single- 
cyhnder pump at Armour Institute of Technology, and curve A based 
on experiments with a 16-in. by 12-in. duplex fire pump at Massachusetts 
Institute of Technology. 



TWO 












Curves of Performance 














/ 












INIarsh Steam Pump 

for 

Varying Speed 

Size of Pump- 12"x T^'x Vl" 

Cap. 216 Gal. per Min, at 100 Strokes 








/ 


/ 


















/ 




6000 












/ 








\ 








/ 


/ 










\ 






































\ 


























/ 










5000 




\ 
























A 












\ 






















/f 










V" 


s 




% 




















/Cj 


■^ 








^ 


^ 


%4 




y 


^ 


















/ 








^ 








^ 4000 






V 
















/ 








^ 






s 




^ 1 


. 




•^ 












; 


/ 


^ 


^ 










i 


15- 


s 


n 






\ 












/4-' 














I 




1 \ 


[ 




\ 


s 






y 


^A 


















w 


.-^ 


oriA 


\ 






\ 


V 


/ 




Y 


















a 




\ 








Xv 


/ 




















^ 






\ 






/ 


!/ 


\ 




















z' 


-10 






y 


\, 


/ 




/ 






\ 


■^ 












^ 


y 




2000 


OAl-l 




> 


\ 


/ 


J 










■^ 


^ 




^ 


y 












/ 


/ 


\ 


/ 
















y 


><: 


^ 


— 


- 










/ 




/ 


S 


r^.. 








>^f" 


















/ 


/ 


/ 








^e. 


^-^ 


















1000 


loo/ 




/ 










^ 


^-^ *. 


^--^ 


.^ 
















/ 








y^ 


X' 
























■""■ 




/ 


/ 




y^ 


^ 
































// 


^ 


^ 




































i^^ 


^'^ 





































30 40 50 60 70 

Single Strokes per Min. 

Fig. 382. 



•^•^ 



90 



lOO 



Fig. 382 gives the details of the performance of a 12-in. by 7J-in. by 12- 
in. Marsh boiler-feed pump at the Armour Institute of Technology. 

The determination of the power consumption of a boiler-feed pump 
is best illustrated by the following example. 

Exmriple 58. A small direct-acting duplex pump uses 150 pounds of 
steam per i.hp-hour. Gauge pressure 150 pounds per square inch; 
feed-water temperature 64 deg. fahr. Required the per cent of rated 
boiler capacity necessary to operate the pump. 



634 STEAM POWER PLANT ENGINEERING 

The head pumped against, 150 pounds per square inch, is equivalent 
to 150 X 2.3 = 345 feet of water. 

The friction through the valves, fittings, and pipe, and the vertical 
distance between suction and feed-water inlet, are assumed to be equiva- 
lent to 20 per cent of the boiler pressure, giving a total head of 150 + 
30 = 180 pounds per square inch, or 414 feet of water. 

A boiler horsepower, taking into consideration leakage losses and 
the steam used by the feed pump, will be equivalent to the evaporation 
of approximately 32 pounds of water per hour from a feed temperature 
of 64 deg. fahr. to steam at 150 pounds gauge. 

The actual work done in pumping 32 pounds of water against a head 
of 414 feet is 

414 X 32 = 13,248 foot-pounds. 

This corresponds to 

13,248 ^ ^^^^ , 

60 X 33,000 ^Q'^^^^^^^^^P^^"^' 

The total heat of one pound of steam above 64 deg. fahr. is 1163 
B.t.u. The heat delivered to the pump per i.hp-hour is 

1163 X 150 = 174,430 B.t.u. 

The amount used by the pump for each boiler horsepower, disregard- 
ing efficiency, is 

174,450 X 0.0067 = 1168 B.t.u. per hour. 

The mechanical efficiency of the average feed pump ranges from 50 
to 85 per cent, depending upon its condition and the number of strokes 
per minute. Assuming it to be 65 per cent, the heat used by the pump 
per hour to deliver 32 pounds of water into the boiler is 
1168 ^0.65 = 1796 B.t.u. 

A boiler horsepower is equivalent to 33,479 B.t.u. per hour. There- 
fore the per cent of boiler output necessary to operate the pump is • 

100 X ^^ = 5.36 per cent. 

If the exhaust steam is used for heating the feed water, the steam con- 
sumption will be 0.73 per cent of the boiler capacity, thus: The weight 
of steam consumed per boiler horsepower-hour 

— — = 1.54 pounds. 

Allowing a 10 per cent loss, the heat in the exhaust available for 
heating the feed water is 

[1150 - (64 - 32)] 0.9 X 1.54 = 1550 B.t.u. 

1796 — 1550 = 246 B.t.u., or the net heat required by the pump per 
hour to deliver 32 pounds of water to the boiler. 



t 



I 



PUMPS 635 

The per cent of boiler output necessary to operate the pump is 

Pump performances are generally given in terms of the foot-pounds 
of work done by the water piston per thousand pounds of dry steam or 
per million B.t.u. consumed by the engine, thus: 
_ Foot-pounds of work done 

Weight of dry steam used ' ^ ^ 

' — Foot-pounds of work done i nnn nnn foKA\ 

Total number of heat units consumed ' ' _ „ 

(See A.S.M.E. code for conducting duty trials of pumping engines, 
Trans. A.S.M.E., Vol. 37, 1915.) 

Example 59. A compound feed pump uses 100 pounds of steam per 
i.hp-hour; indicated horsepower, 48; capacity, 400 gallons per minute; 
temperature of water, 200 deg. fahr.; total head pumped against, 
175 pounds per square inch; steam pressure, 100 pounds gauge; moisture 
in the steam, 3 per cent. Required the duty on the dry steam and on 
the heat-unit basis. 

175 pounds per square inch is equivalent to 175 X 2.4 = 420 feet of 
water at 200 deg. fahr. 

Weight of 400 gallons of water at 200 deg. fahr. = 400 X 8.03 = 3212 
pounds. 

Work done per minute = 3212 X 420 = 1,329,040 foot-pounds. 

Weight of dry steam supplied per minute 

100 X 48 ^ „ ^_ „_ „ , 

= — — X 0.97 = /7.6 pounds. 

B.t.u. supphed per minute 

= ^^^^ ^^ (0.97 X 879.8 + 309 - 200 -f 32) = 79,552. 
Duty per thousand pounds of dry steam 

= ^>^^^^-Q^Q X 1000 = 17,384,150 foot-pounds. 
Duty per million B.t.u. 

= ^79 552^ X 1,000,000 = 16,958,000 foot-pounds. 

Table 111 may be used in approximating the duty, thus: 

The mechanical efficiency of the pump in the preceding problem is 

j,^ . P.hp. 1,349,040 ^ ^^ 

Efficiency = ^^ = 33 qqq ^ ^3 = 85 per cent. 

At the intersection of vertical column ''85" and horizontal column 
100" of Table 111, we find 16.82 millions. See, also, Table 79. 



636 



STEAM POWER PLANT ENGINEERING 



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m 
•^ O 

Q 



r/j 


c 


P 


P 


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PUMPS 



637 



I 



t 



Tables 112 and 113 give the maximum theoretical height to which 
pumps may lift water by suction at different temperatures. In prac- 
tice these figures cannot be realized. It is customary to have the water 
gravitate to the pump for all temperatures over 120 deg. fahr. 



TABLE 112. 

MAXIMUM HEIGHTS TO WHICH PUMPS CAN RAISE WATER BY SUCTION. 
(Temperature of Water 40 Deg. Fahr.; Barometer 29.92.) 



Vacuum in 

Suction Pipe, 

Inches of 


Theoretical 

Lift. 


Probable 

Actual 

Lift. 


Vacuum in 

Suction Pipe, 

Inches of 


Theoretical 
Lift. 


Probable 

Actual 

Lift. 


Mercury. 




Mercury. 






Feet. 


Feet. 




Feet. 


Feet. 


1 


1.1 


0.9 


16 


18.0 


14.4 


2 


2.2 


1.8 


17 


19.1 


15.3 


3 - 


3.3 


2.7 


18 


20.2 


16.1 


4 


4.5 


3.6 


19 


21.4 


17.1 


5 


5.6 


4.5 


20 


22.5 


18.0 


6 


6.7 


5.4 


21 


23.7 


18.9 


7 


7.9 


6.3 


22 


24.8 


19.8 


8 


9.0 


7.2 


23 


25.9 


20.7 


9 


10.1 


8.1 


24 


27.0 


21.6 


10 


11.3 


9.0 


25 


28.2 


22.7 


11 


12.4 


9.9 


26 


29.3 


23.9 


12 


13.5 


10.8 


* 27 


30.4 


24.3 


13 


14.6 


11.7 


28 


31.6 


25.2 


14 


15.8 


12.6 


29 


32.7 


26.1 


15 


16.9 


13.5 


t 29.68 


33.6 









* Vacua greater than 27 inches are practically unobtainable in pumping practice except in 
connection with condensers. 

t Maximum theoretical vacuum obtainable with water at 40 degrees F. and barometer of 
29.92 inches. 



TABLE 113. 

MAXIMUM THEORETICAL HEIGHT TO WHICH A PUMP CAN LIFT WATER BY 

SUCTION AT DIFFERENT TEMPERATURES. 

(Barometer 29.92.) 



Temperature of 


Maximum 


Temperature of Feed 


Maximum 


Feed Water. 


Theoretical Lift. 


Water. 


Theoretical Lift. 


Deg. fahr. 


Feet. 


Deg. fahr. 


Feet. 


40 


33.6 


130 


29.2 


50 


33.5 


140 


27.8 


60 


33.4 


150 


25.4 


70 


33.1 


160 


23.5 


80 


32.8 


170 


20.3 


90 


32.4 


180 


16.7 


100 


31.9 


190 


12.8 


110 


31.3 


200 


7.6 


120 


30.3 


210 


1.3 



638 STEAM POWER PLANT ENGINEERING 

^96. Size of Boiler-feed Pump. — Reciprocating Piston Type. — Let 

D = diameter of water cylinder, inches. 

d = diameter of the steam cyhnder, inches. 

L = length of stroke, inches. 

N = number of working strokes per minute. 

H = head in feet between suction and boiler water level. 

R = resistance in pounds per square inch between suction level 
and boiler water level due to valves, pipes, and fittings. 

p = boiler pressure, pounds per square inch. 

>S = ratio of the water actually delivered to the piston displace- 
ment. 

W = weight of water delivered, pounds per hour. 

/ = indicated horsepower of the pump at maximum capacity. 

E = mechanical efficiency of the pump, taken as the ratio of the 
water horsepower at the discharge opening to the indi- 
cated horsepower of the pump, steam end. 



Then 



Tf=^.-5!.Mx60x 62.5 X ^ = 1.7 D^LNS. (255) 
4 144 ]2 ^ ^ 



D = 0.77 



V/^- (256) 

V LNS 



D^V_ 



+ /^ + 0.433^^ (257) 



Ey 

_ TF (p + ig + 0.433 ff) 2.3 
^ " 33,000 X 60 X ^ ' ^^^^^ 

In average practice the piston or plunger displacement is made about 
twice the capacity found by calculation from the maximum amount of 
water required for the engine, to allow for leakage, steam consumption of 
the auxiliaries and blowing off. 

For pumps with strokes of 12 inches or over, the speed of the plunger 
or piston is usually limited to 100 feet per minute as a maximum to in- 
jure smooth running. For shorter strokes a lower limit should be used. 
The maximum number of strokes ranges from 100 for strokes over 12 
inches in length to 200 for strokes under 5 inches. Boiler-feed pumps 
should be designed to give the desired capacity at about one-half the 
maximum number of strokes or less. 

Pump slip varies from 2 to 40 per cent, depending upon the condition 
of the piston and valves and the number of strokes. An average value 
for piston and plunger pumps in first-class condition is 8 per cent when 
operating at rated capacity, but it is wise to allow a much larger figure, 
say 20 per cent, for leakage caused by wear. 



PUMPS 639 

The area of the steam cylinder is made from 2 to 2.5 times that of 
the water end to allow for the various friction losses and the drop in 
pressure between the pump throttle and the boiler. The total head 
pumped against includes the suction lift, the friction of valves and 
fittings, the distance between the suction inlet and the boiler level and 
the boiler pressure. The excess head varies in practice from 15 to 40 
per cent of the boiler pressure; an average figure is 25 per cent. In 
allowing for the drop in steam pressure between boiler and pump a 
liberal figure is 25 per cent. 

The apphcation of equations (255) to (258), including the practical 
considerations stated above, is best illustrated by a specific example. 

Example 60. Determine the size of direct-acting single-cylinder feed 
pump necessary to supply water to 1000 horsepower of boilers operating 
at rated capacity. Gauge pressure 100 pounds per square inch; feed- 
water temperature 150 deg. fahr. 

One horsepower is equivalent to the evaporation of 34.5 pounds of 
water from and at 212 deg. fahr.; but the pump is usually designed to 
supply about twice the required amount of water. 

Thus W = 62,400 (under the given conditions). 
S = 0.8 (by assumption). 
LN = 1200 (on the basis of 100 feet per minute). 

Substitute these values in (256) : 



D = 0.77 y — 5?477r^ = 6.2 inches, — call it 6 inches, 



1200 X 0.8 

since the assumptions have been very liberal. 
Assume (0.433 H + R) = 0.25 p Siud E = 0.65. 
Substitute these values in (257) : 



Vo 



100 + 25 



.65 X 100 
= 8.35, — call jt 8.5 inches. 

Allowing 100 strokes per minute the length of the stroke must be 
L = 1200 ^ 100 = 12 inches. 

The dimensions of the pump are 8J-in. by 6-in. by 12-ii.. 
The indicated horsepower at maximum load may be obtained by 
substituting the proper values in (258), thus: 

62,400(100 + 25)2.3 



/ = 



33,000 X 60 X 0.65 
13.9 i.hp. 



297. Steam-pump Governors. — Fig. 383 shows a section through a 
Fisher pump governor, illustrating a device for maintaining a practically 
constant pressure in the discharge pipe irrespective of the quantity of 



640 



STEAM POWER PLANT ENGINEERING 



water flowing. It embodies a pressure-reducing valve in the steam 
supply pipe of the pump, actuated by the slight variations in water 
pressure. When the demand for water increases, the pressure in the 
discharge pipe tends to decrease, and this drop in pressure (transmitted 
g to the pump governor by suitable piping) causes 

more steam to be admitted, which increases the 
speed of the pump. The governor is connected 
to the steam inlet of the pump at B and the 
steam enters at A. Double-balanced valve C 
regulates the supply of steam to the cyUnder by 
the amount it is raised from the seat. The valve 
is held open by spring G, the compression of 
which may be regulated by hand wheel K. The 
water pressure from the discharge pipe acts on 
piston F and tends to overcome the resistance 
of the spring. The difference in pressure between 
the water and the spring determines the position 
of valve C. 

Piston rod H is pinned to sleeve / and valve 
stem L screwed into this sleeve by means of 
hand wheel K. Hence, during ordinary opera- 
tion, the piston, piston-rod sleeve, valve stem, 
and valve act as a single unit. By turning the 
hand wheel K, valve stem L will screw into 
sleeve / and the tension on the spring will 
be increased. Hand wheel J serves as a lock 
Fig. 383. Fisher Pump j^^^ ^Jid prevents K from turning during nor- 
Governor. ^^j operation. 

298. Feed-water Regulators. — The water level in the boiler should 
be kept as nearly constant as possible, and this necessitates considerable 
attention on the part of the fireman, especially with fluctuating loads. 
There are a number of devices on the market which are designed to 
automatically maintain a constant level, and in many small plants 
where the duties of the fireman are numerous such devices in connec- 
tion with high and low water alarms are of considerable assistance. 
Their action, however, is not always positive on account of wear or 
sticking of parts, and engineers as a rule prefer to rely upon hand regu- 
lation. 

Fig. 384 shows a section through a Kitts feed-water regulator, con- 
sisting of two parts, the chamber F and the regulating valve V. The 
float chamber is connected to the boiler or water column at and E, 
and the regulating valve to the feed main at R and to the boiler feed 




PUMPS 



641 



pipe at W. When the water in the boiler falls below the mean level, the 
weight B overcomes the counterweight G and closes needle valve L by 
means of compound levers. At the same time an extension on valve L 
Kfts spring A and opens exhaust valve D. This removes the steam 
pressure from the top of diaphragm C, in the regulating valve, through 
the agency of pipe K. The pressure from thr pump raises the disk T 
and water flows into the boiler until the watei rises to the mean level. 




■ — Mean- - 




Fig. 384. Kitts Feed-water Regulator. 
(Counterbalanced-weight Type.) 



Fig. 385. Rowe Feed-water 
Regulator. (Float Type.) 



When weight B becomes submerged its weight is overcome by counter- 
weight G, valve L is opened and exhaust valve D is closed. This ad- 
mits steam pressure to the diaphragm C and forces disk T to its seat, 
cutting off the supply of water to the boiler. 

The Rowe feed-water regulator, Fig. 385, depends for its operation 
on a famiUar float-controlled valve mechanism. The vessel A is con- 
nected to the boiler above and below the water line, and the float C, 
following the water level up and down, actuates a balanced valve in 
accordance with the boiler-feed requirements. When this apparatus 
is used to regulate the feed of a single boiler the opening G in the valve 
chamber is connected to the steam space of the boiler and the outlet H 
is carried to the steam inlet of the feed-water pump. When the water 
level is normal the float closes the valve L and thereby cuts off the sup- 



642 



STEAM POWER PLANT ENGINEERING 



ply of steam to the pump cylinders. Communication between chambers 

A and R is prevented by means of a 
diaphragm M. When the water level 
falls below normal the float pulls the 
valve down, opening the way for steam 
to pass from the inlet G to the outlet 
H and thence to the pump. When the 
regulator is used to control a battery 
of boilers the pump discharge delivers 
into the inlet G and the water passes 
through H to the boiler-feed main. 
Should the water level fall beyond a 
predetermined limit by reason of any 
accidental discontinuance of the water 
supply which the apparatus cannot 
correct, the float would open the valve 
F of the alarm whistle mounted on 
the top of the main vessel. 

Fig. 386 gives the general details of 
the '^S-C" feed-water regulator which 
differs from the types just described 
in the manner of actuating the water 
regulating valve. A small copper ves- 
sel, A, partly filled with water, is in 
communication with diaphragm I 
The water in vessel A is independent 
U, projects into chamber 




Fig. 386. 
lator. 



''S-C" Feed-water Regu- 
(Thermo-pressure Type.) 



through the medium of tube F. 

of the boiler supply. A small copper U-tube 




Fig. 387. Copes Feed-water Regulator. (Thermo-expansion Type.) 



A, as indicated. When the water in the boiler is at its highest level 
the U-tube is filled with water and the pump regulator valve V is not 



PUMPS 



643 





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Knowles High Speed Electric Pump 

Direct connected to 

M P 6-100 H.P. -280-220 V.Form L 

Load and Efficiency 
































































































































































































































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Gauge Pressure at Valve (Lb,) 
Fig. 388. 



644 



STEAM POWER PLANT ENGINEERING 



feeding. As the level of the water in the boiler drops, the water recedes 
from the outer surface of the U-tube, and the upper branch of the tube 
is surrounded with steam. The steam causes the water in the vessel A 
to boil, and the pressure generated is transmitted through pipe F to 
diaphragm I, thereby opening controlUnq: valve K. Wheel J permits of 
hand control. Regulators of this type installed in [the power plant of 
the Armour Institute of Technology are giving excellent service. 

The Copes feed-water regulator. Fig. 387, depends for its operation 
upon the expansion and contraction of an inclined tube. As illus- 
trated, this inclined tube is so placed that it contains steam when the 
water in the boiler is at its lowest level. As the water gradually rises 
in the boiler it rises in the tube also. When the level of the water is as 

shown in the illustration the part 
of the tube filled with steam is at 
boiler pressure and temperature and 
that part containing water is at a 
lower temperature. When an in- 
creased load comes on there is a 
slight drop in steam pressure, ac- 
companied by a more rapid libera- 
tion of steam from the entire body 
of water within the boiler, causing 
a corresponding increase in its vol- 
ume and a rise in the boiler water 
level. This at once raises the level 
of the water in the expansion tube 
slightly, increases the amount of the 
tube submerged in the water and 
decreases the tube temperature, 
causing it to shorten. Since this 
tube is connected by a simple system of levers to a balanced valve in 
the feed Une, shortening of the tube causes the valve to close, so that in- 
crease in water level results in a decrease in the rate of feed. 

299. Power Pumps. — Piston Type. — Piston pumps, geared, belted, 
or direct connected to electric motors, gas engines, and water motors, are 
used chiefly where steam power is not available. Their general utility 
is evidenced by the rapidly increasing number installed in situations 
formerly occupied by the direct-acting steam pump. The efficiency of 
this type of pump depends in a large measure upon the character of 
the driving motor and the efficiency of the transmitting mechanism. 
High-speed power pumps direct connected to electric motors give 
efficiencies from Hne to water horsepower as high as 83 per cent, while 




Fig. 389. 



A Typical Geared Triplex 
Pump. 



PUMPS 



645 



the low-speed geared type seldom exceed 70 per cent. The curves in 
Fig. 388 give the performance of a direct-connected triplex pump, 
and those in Fig. 390 the performance of a triplex pump geared to an 
electric motor. Both of these performances are exceptionally good and 
are considerably above the average. 

For a General Treatise on the Design and Operation of Pumping Machinery con- 
sult "Pumping Machinery," by A. M. Greene; John Wiley & Sons, 1911. 

100 



U 50 



I 



25 









. 


^_Motor_ 
Pump 






md Gearin 


r 












^ 


^' 


Set 




















Head Cons 


ant, 350 Ft. 



















100 



84 a 
WS 50 

^^ 

-8 

^ 25 



200 400 600 800 1000 

Discharge, Gallons Per Minute 
Head Constant. Speed Variable. 



1200 



1400 









Motor 






- 




pu«oP_^ 


id Gearing 
















—- 


Set 














y^ 






10 X 

( 


2 Triplex P 
5 II. P. Mot( 


imp 

r 










M( 
Capacit 


tor R.P.M. 
earing 10 to 
y 1250 Gal. p 


r 

;r Min. 



50 75 100 125 

Total Head, Lb. Per Sq. In. Gauge 
Speed Constant. Head Variable. 



150 



175 



Fig. 390. Performance of a 65-horsepower, Motor-driven Triplex Pump. 

Geared Type. 

300. Injectors. — As a boiler feeder the injector is an efficient and 
convenient device, cheap and compact, with no moving parts, delivers 
hot water to the boiler without preheating, and has no exhaust steam 



646 



STEAM POWER PLANT ENGINEERING 



to be disposed of. Its adoption in locomotives is practically universal, 
but in stationary practice it is limited to small boilers or single boilers 
or as a reserve feeder in connection with pumps. The objections to 
an injector are its inabiUty to handle hot water, the difficulty of main- 
taining a continuous flow under extreme variation of load, and the un- 
certainty of operation under certain conditions. Fig. 391 illustrates 
the simplest form of single-tube injector. Boiler steam is admitted at 
A and, flowing through nozzle and combining tube to the atmosphere 
through G, partially exhausts the air from pipe B, thereby causing the 
water to rise until it comes in contact with the steam. The steam 
emerging from nozzle C at high velocity condenses on meeting the water 



Steam Supply 




Boiler 



Fig. 391. Elementary Steam Injector. 

and imparts considerable momentum to it. The energy in the rapidly 
moving mass is sufficient to carry it across opening 0, lift check H 
from its seat and force it into the boiler. The steam then ceases to 
escape at G. 

301. Positive Injectors. — Fig. 392 shows a section through a Han- 
cock injector, illustrating the principles of the double-tube positive 
type. Its operation is as follows : Overflow valves D and F are opened 
and steam is admitted, which at first passes freely through the over- 
flow to the atmosphere and in so doing exhausts the air from the suction 
pipe. This causes the feed water to rise until it meets the jet of steam 
and the two are forced through the overflow. As soon as water appears 
at the overflow, valve D is closed, valve C partially opened, and valve 
F closed. This admits steam through the forcing jet W and, the over- 
flow valves being closed, the water is fed into the boiler. In case the 
action is interrupted for any reason it is necessary to restart it by hand. 

The chief advantage of the double-tube positive type lies in its 
abihty to lift water to a greater height and to handle hotter water than 
the single-tube. Its range in pressure is also greater, that is, it will 
start with a lower steam pressure and discharge against a higher back 
pressure. Double-tube injectors are used almost exclusively in loco- 
motive work. 



PUMPS 



647 



302. Automatic Injectors. — Fig. 393 shows a section through the 
Penberthy injector. Its operation is as follows: Steam enters at the 
top connection and blows through suction tube c into the combining 
tube d and into chamber g, from which it passes through overflow valve 
n to the overflow m. When water is drawn in from the suction intake 
and begins to discharge at the overflow, the resulting condensation of 
the steam creates a partial vacuum above the movable ring h and the 
latter is forced against the end of tube c, cutting off the direct flow 
of water to the overflow. The water then passes into the boiler. Spill 



steam 




Overflow 

Fig. 392. Hancock Double- 
tube Injector. 




Fig. 393. Penberthy Automatic 
Injector. 



holes i, i, i are for the purpose of relieving the excess of water until 
communication with the boiler has been estabUshed. The action of 
opening and closing the overflow is entirely automatic. Where the 
conditions are not too extreme the automatic injector is to be preferred 
for stationary work because of its restarting features. It is also used on 
traction, logging, and road engines, where its certainty of action and 
special adaptability render it invaluable for the rough work to which 
such machines are subjected. 

Injectors, Theory of: Trans. A.S.M.E., 10-339; Sibley Jour., Dec, 1897, p. 101; 
Power, May, 1901, p. 23; Thermodynamics of the Steam Engine, Peabody, Chap. 
IX; Theory of the Steam Injector, Kneass. 

Injectors, General Description: Eng;r. U. S., Oct. 1, 1907, Nov. 15, 1907, July 15, 
1904, p. 501, Feb. 2, 1903, p. 151; Power, Aug. 1906, p. 478; Engr., Lond., March 
10, 1905, p. 244; Engineering, Aug. 30, 1895, p. 281. 



648 



STEAM POWER PLANT ENGINEERING 



ft to 

2 ^ 



c 


) 




























\ 






Constant Discharge Pressure 

20 Lb. Per Sq.In. 

Constant Suction Temp. 

55 Deg.Fah. 














\J 


) 


















<^ 




























^ 


\^ 






























v^^C 


) 






























"^ 


































) 


"--s 


3 



65 



70 



75 



90 



95 



Initial Gauge Pressure,Lb.Per Sq.In. 
Fig. 394. Performance of an Automatic Injector with Varying Initial Pressure. 



20 



2 a 



17 



16 



,15 



H 





































Constant Initial Pressure 

70 Lb. Per Sq. In. 

Constant Discharge Pressure 

70 Lb. Per Sq. In. 








c 


r^" 






















( 


D"^ — 
































-^ 




-Q^ 






c 




























" 





55 



65 



75 85 95 

Temperature of Suction, Deg. Fah. 



105 



115 



Fig. 395. Performance of an Automatic Injector with Varying Suction Temperature. 



^1 
si 



A" 



a> It 



16 





Constant Initial Pressure,70 Lb. Per Sq. In. 
Constant Suction Temperature, 56 Deg. Fah. 












c 










i 


2 






















1 


) 
















) 

































80 



30 40 50 60 

Discharge Pressure, Lb. Per Sq, In. Gauge 



80 



Fig. 396. Performance of an Automatic Injector with Varying Discharge Pressure. 



PUMPS 



649 



303. Performance of Injectors. — The performance of an injector may 
be very closely determined from the equation 

xr -\- q — t + 32 (Kneass, ''Theory of the .^_^v 

^ = r:ri Injector, " p. 83), ^^^^'^ 

in which 

w = pounds of water delivered per pound of steam suppUed, 
X = quaUty of the steam supplied, 
r = heat of vaporization, 
q = heat of the liquid, 
t = temperature of the discharge water, 
to = temperature of the suction water. 

Figs. 394, 395, and 396 give the performance of a Desmond auto- 
matic injector as tested at the Armour Institute of Technology. The 
results check very closely with those calculated from above equation. 
Referring to Fig. 394 it will be seen that the weight of water deUvered 
per pound of steam decreases as the initial pressure is increased, all 
other factors remaining the same. From Fig. 395 it will be noted that 
the weight of water deUvered per pound of steam decreases as the tem- 
perature of suction supply is increased up to a point where the injector 
''breaks" or becomes inoperative. This critical temperature varies 
with the different types of injectors, being highest for the double-tube 



TABLE 114. 

RANGE IN WORKING PRESSURES. 
Standard " Metropolitan " Steam Injectors. 





Automatic. 


Suction 
Temperature, 




s 


jction Head, Feet. 




Deg. Fahr. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 
100 


25 to 150 

26 to 120 


30 to 130 
33 to 100 


42 to 110 
55" to 80 


55 to 85 


20 to 160 
25 to 125 


120 




26 to 85 


140 
























Double Tube. 


Suction 
Temperature, 


Suction Head, Feet. 


Deg. Fahr. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 
100 
120 


14 to 250 

15 to 210 
20 to 185 
20 to 120 


23 to 220 

26 to 160 
30 to 120 
35 to 70 


27 to 175 
37 to 120 
42 to 75 


42 to 135 
46 to 70 


14 to 250 

15 to 210 
20 to 185 


140 




20 to 120 











650 STEAM POWER PLANT ENGINEERING 

type, but seldom exceeds 160 deg. fahr. Fig. 396 shows that the weight 
of water dehvered per pound of steam is practically constant for all 
discharge pressures within the limits of the apparatus. 

Table 114 gives the range of working steam pressures for standard 
*' Metropolitan" injectors with varying suction heads and temperatures, 
and, though strictly applicable to this particular type only, is charac- 
teristic of all makes. 

In selecting an injector the following information is desirable for 
best results: 

1. The lowest and highest steam pressure carried. 

2. The temperature of the water supply. 

3. The source of water supply, whether the injector is used as a lifter 
or non-hfter. 

4. The general service, such as character^ of the water used, whether 
the injector is subject to severe jars, etc. 

304. Injector vs. Steam Pump as a Boiler Feeder. — From a purely 
thermodynamic standpoint the efficiency of an injector is nearly per- 
fect, since the heat drawn from the boiler is returned to the boiler again, 
less a shght radiation loss. As a pump, however, the injector is very 
inefficient and requires more fuel for its operation than very wasteful 
feed pumps. This is best illustrated by an example: 

Example 61. a^Compare the heat consumption of a high-grade injector 
with that of an ordinary duplex boiler feed pump when feeding water 
to a boiler. Make all necessary assumptions. An injector of modern 
construction will deliver say 15 pounds of water to the boiler per pound 
of steam supplied, with delivery temperature of 150 deg. fahr. This 
corresponds to a heat consumption of 71.3 B.t.u. per pound of water 
delivered, thus: 

With initial pressure of 115 pounds absolute, 

H = 1188.8. 

Heat of the water delivered to the boiler, 

150 - 32 = 118 B.t.u. above 32 deg. fahr. 

Heat of 1 pound of steam above a feed temperature of 150 deg. fahr., 

1188.8 = 118 = 1070.8 B.t.u. 

Heat required to deliver 1 pound of water to the boiler, 

1070.8 



15 



71.3 B.t.u. 



A simple direct-acting duplex pump consumes say 200 pounds steam 
per i.hp-hour. Assume the extreme case where the exhaust steam will 
not be used for heating the feed water and the latter is fed into the 
boiler at 60 deg. fahr. 



PUMPS 651 

The heat supplied to the pump per i.hp-hour, 

200 11188.8 - (60 - 32)j = 232,160 B.t.u. 

Assuming the low mechanical efficiency of 50 per cent, the heat re- 
quired to develop one horsepower at the water end will be 

232,160 -^ 0.50 = 464,320 B.t.u. per hour. 

Since the steam pressure is 100 pounds gauge, the equivalent head 
of water at 60 deg. fahr. is 

2.3 X 100 = 230 feet. 

Assume the friction in the feed pipe, the resistance of valves, etc., to 
be 30 per cent of the boiler pressure; the total head pumped against 
will be 

230 + 69 = 299, say 300 feet, 

1 horsepower-hour = 1,980,000 foot-pounds per hour, 

1,980,000 



300 



= 6600 pounds ; 



that is, 1 horsepower at the pump will deliver 6600 pounds of water per 
hour to the boiler against a head of 300 feet. 

The heat consumption per pound of water delivered, 

464,320 _ 7„ , R . „ 

"66or - ^°-^ ^•*"- 

If the feed water is heated to say 210 deg. fahr. by the exhaust steam 
from the pump, the heat consumption will be 63.7 B.t.u. as against 70.3 
without the heater. 

Thus even in this extreme case of poor steam-pump performance 
the heat consumption lies in favor of the pump. With the better 
grades of pumps this disparity is considerably greater, and decidedly 
so if the exhaust steam is used to preheat the feed water. For inter- 
mittent operation the condensation losses in the pump may more than 
offset this gain. Other conditions, however, such as compactness, 
low first cost, and ease of operation are oftentimes considerations and 
the heat consumption is of minor importance. 

305. Vacuum Pumps. — The different types of vacuum pumps em- 
ployed in steam power plant practice may be divided into four general 
classes : 

1. Wet-air pumps. 

2. Tail pumps. 

3. Dry-air pumps. 

4. Condensate pumps. 

(1) Wet-air pumps are for the purpose of withdrawing water and non- 
condensable gases from apparatus under less than atmospheric pressure. 
Standard low level jet-condenser wet-air pumps handle simultaneously 



652 



STEAM POWER PLANT ENGINEERING 



the circulating water, condensate, and all entrained air and are, in fact, 
a combination of circulating pump and vacuum pump. Surface con- 
denser wet-air pumps deal with the condensate and its air entrainment. 
Wet-air pumps may be of the reciprocating, centrifugal, rotary jet, 
rotary positive displacement and steam jet type. 

(2) The terms ''wet-vacuum pump," ''wet-air pump," and "tail 
pump " are often used synonymously but in order to differentiate between 
pumps handling injection water, condensate and air from those deaUng 
only with the injection water and condensate the term "wet-air pump" 
has been applied to the former and "tail pump" to the latter. 

(3) Dry-air pumps are for the purpose of withdrawing the non- 
condensable gas from apparatus under a vacuum and discharging it 
against atmospheric or greater pressure. They are to all intents and 
purposes, air compressors. The term "dry air" is a misnomer since the 
gases exhausted are almost invariably saturated with water vapor. 

These pumps may be of the recipro- 
cating, rotary, positive displace- 
ment, hydro-centrifugal and steam 
jet types- 

(4) Condensate pumps are for the 
purpose of withdrawing condensed 
steam from surface condensers and 
are usually of the] reciprocating, ro- 
tative or centrifugal types. 

306. Wet-air Pumps for Jet Con- 
densers. — Fig. 397 shows a section 
of the cylinder of a Dean twin- 
cyHnder wet-air pump as applied to 
a standard low-level jet condenser 
and illustrative of the reciprocat- 
ing type. There are three sets of 
valves, the suction or foot valves A, 
A, the lifting or bucket valves B, B, 
and the head or discharge valves C, C. On the upward stroke of 
the piston or bucket a partial vacuum is formed in the chamber be- 
tween the bucket and the lower head, causing the water and air in 
the bottom of the barrel to lift the foot valves A, A from their seats 
and flow into the cylinder. On the downward stroke the foot 
valves A, A close and water and air are entrapped in chamber R be- 
tween the lower head and the bucket. As the bucket descends, the 
pressure of air in the cylinder lifts the bucket valves B, B from their 
seats and permits the air and water to escape to the upper portion S 




Fig. 397. Dean Air Pump. 



PUMPS 



653 



of the cylinder between the head plate and the bucket. On the next 
upward stroke the water and air are forced through the discharge valves 



Exhaust Steam 




Circulating 
Water 



Fig. 398. Rees ''Roturbo" Jet Condenser. 



C, C into the hot well. This discharge of water and air from the top 
compartment is simultaneous with influx of water and air in the lower 
chamber. 

Fig. 398 shows a vertical section and sec- 
tional end elevation of a Rees Roturbo ro- 
tary jet condenser illustrating an adaptation 
of the rotary-jet pump as a jet condenser. 
This pump is a development of a special 
type of centrifugal pump the unique feature 
of which is the employment of a revolving 
pressure chamber. The hollow impeller, 
Fig. 399, lifts the circulating water in much 
the same manner as in any centrifugal 
pump. The space between the periphery of 
the impeller and the inner circumference of 
the fan wheel forms the mixing chamber in 
which the exhaust steam is brought into 
contact with radial jets of water. The fan 
wheel itself acts as an ejector and exhausts 
the mixture of circulating water and vapor. The operation is as follows : 
circulating water is drawn through the suction pipe into the revolving 
pressure chamber, on the periphery of which nozzles are arranged as 
shown in Fig. 399, and is forced through the nozzles in radiating jets 




Fig. 399. Impeller for Rees 
' ' Roturbo ' ' Jet-condenser 
Pump. 



654 



STEAM POWER PLANT ENGINEERING 



which are arranged to impinge in pairs. The water jets, which are made 

fan shaped and subdivided into a fine spray, are projected in fines 

radiating from the shaft (but stifi rotating as a whole with the impefier) 

across a space into which the exhaust steam blows. The circulating 

water leaving the nozzles, condensate, and air entrainment are picked 

up by the blades of the fan and discharged through a volute guide 

chamber to the hot well. 

The Connersville jet condenser is a typical example of an application 

of a rotary positive-displacement wet-air pump. In this device the 

circulating water, condensate, and air entrainment are handled by a 

Connersville cycloidal 3-lobe type rotary pump. (A cross section 

through a typical 2-lobe cycloidal pump is shown in Fig. 424.) 

The steam-jet type of wet-air pump is exemplified in the ejector 

condenser. See paragraph 237. 

306a. Wet-air Pumps for Surface Condensers. — These pumps exhaust 

the condensate and air entrainment from surface condensers. The 

vacuum pumps of a steam heating system come also under this head. 

The Edwards air pump, Fig. 400, is a 

typical example of a wet-air pump of the 

reciprocating type. Referring to Fig. 400, 

the condensed steam flows continuously by 

gravity from the condenser into the base 

of the pump through passage A and annular 

space B. As the piston C descends it forces 

the water from the lower part of the casing 

F into the cyfinder proper through the ports 

P, P. On the upward stroke the ports in 

the piston are closed and the air and water 

discharged through head valves D and ex- 
Fig. 400. Edwards Air Pump, j^^^^ p^^ ^ ^^ ^j^^ j^^^ ^^1, The seats of 

valves D are constructed with a rib between each valve and a lip 
around the outer edge, so that each valve is water-sealed independently 
of the others. In ordinary air pumps the clearance between the bucket 
and head valve seat is necessarily large, due to the space occupied by 
the bucket valves and the ribs on the under side of the valve seating. 
This clearance space reduces the capacity of the pump, since the air 
above the bucket must be compressed above atmospheric pressure be- 
fore it can be discharged, and on the return stroke will expand and 
occupy a space which should be available for a fresh supply of air from 
the condenser. In the Edwards air pump the clearance space is re- 
duced to a minimum, since there are no bucket valves to limit it. The 
absence of suction or foot valves still further increases the capacity of 




PUMPS 



655 



the pump for similar reasons. These pumps are arranged either single, 
double, or triplex; steam, electric, or belt driven; slow or high speed. 
Fig. 401 shows a partial axial and an end section through a C. H. 
Wheeler & Co.'s high- vacuum '^Rotrex" pump. This pump is of the 
wet-vacuum type and handles both air and water of condensation but 
it is also adapted for dry air purposes. The apparatus consists of a 
cylindrical casing and a rotor mounted eccentrically on the shaft. This 
shaft is carried in outboard ring oil bearings which are entirely inde- 
pendent of the stuffing boxes. The division between the suction and 
discharge space in the pump cylinder is maintained by a radius cam 
carried on a shaft independent of the stuffing boxes. This cam is 
operated from the rotor shaft by a lever and crank on the outside of 
the casing. The clearance spaces are water sealed. The discharge 




High- vacuum ''Rotrex" Pump. 



valves are of the Gutermuth type. Pump speed 200 to 300 r.p.m. 
The manufacturers guarantee that on dead-end test a vacuum may be 
obtained within one half inch of the barometer, and within one inch of 
the barometer under operating conditions. 

307. Size of Wet-air Pumps. — Since the wet-air pump for jet con- 
denser must deal with the mixture of injection water, condensate, and 
all air entrainment, the problem of design is essentially that of determin- 
ing the volume of mixture to be withdrawn under condenser pressures 
and temperatures. The volume of injection water and condensate for 
a given set of conditions may be readily calculated, but the volume of air 
entrained with the injection water and condensate and that introduced 
by leakage is an unknown quantity and can only be estimated. The 
amount of air mechanically mixed with the injection may vary from 1 
to 5 per cent by volume at atmospheric pressure and temperature. The 
amount of air in feed water varies from less than 1 per cent by volume, 



656 STEAM POWER PLANT ENGINEERING 

if the heater is of the open type, to 5 per cent or more if the heater is 
of the closed type and raw water is fed directly into the heater. Air 
leakage is an unknown quantity varying within wide limits and is 
dependent upon the tightness of joints, stuffing boxes and the Hke. A 
very liberal factor is usually allowed in estimating total air entrainment, 
an average figure being about 10 per cent by volume of the circulating 
water for the combined air and wet-vacuum pump for jet condensers 
and 10 per cent by volume of the feed water for surface condensers. 
Let Q = total volume of air and water in cubic feet per hour to be 
handled by the pump, 
V = volume of cooling water in cubic feet per hour, 
V = volume of condensed steam in cubic feet per hour, 
Va = volume of air at pressure pa and temperature ta, 
ta = temperature of the air entering the condenser, deg. fahr., 
ti = temperature of the discharge water, deg. fahr., 
U = initial temperature of the cooling water, deg. fahr., 
Pa = atmospheric pressure, pounds per square inch, 
Pc = total pressure in the condenser, pounds per square inch, 
-pv = pressure of aqueous vapor at temperature ^2, 
then {V + v) = volume of water to be pumped from the condenser 
per hour. 

The air entering the condenser will be increased in volume on account 
of the reduction in pressure and the increase in temperature. If Va is 
the original volume under pressure pa and temperature ta the final 
volume on entering the condenser is 

Final volume = Va —^^- X !M^' (260) 

Pc - Pv ta + 460 

and the total volume to be exhausted per hour by the pump is 

^ = ^ + ^ + %-:^.x;7tS- (^-) 

Example 62. Estimate the piston displacement of a wet-air pump 
suitable for average reciprocating engine practice. 

Under average conditions of reciprocating-engine practice the hot- 
well temperature is about 110 deg. fahr. and the absolute back pressure 
4 inches of mercury. Assuming 70 deg. fahr. as the initial temperature 
of the circulating water and allowing 10 per cent as the air entrainment, 

Pa = 29.92 ^ = 70 V = 0.04 V 

Pc = 4: h= 110 Va = 0.1 V. 

py = 2.59 ta = to = 70 

Substitute these values in (261) 

r^ Tr ■ nn. Tr ■ niTr 29.92 110 + 460 

= 3.3 K 



PUMPS 657 

Average practice gives 3 V as the pump displacement per hour for a 
single-acting pump and 3,5 V for a double-acting pump, the cylinders 
being ordinarih' proportioned on a piston velocity of 50 feet per minute 
at rated capacity. 

Wet-air pumps are usually independently driven, making it possible 
to vary the speed of the pump irrespective of the engine speed and to 
create a vacuum before starting the engine. Occasionally, however, 
when the load is constant, as in pumping-engine practice, the pump 
may be driven by the main engine. 

The combined air, condensate and circulating pump (with the ex- 
ception of pumps of the Rees ''roturbo jet" type) is not adapted for 
high-vacuum work on account of the enormous increase in air volume at 
very low pressures. With cold injection water and a good air-tight 
condensing system vacua as high as 2 inches absolute are possible 
with the standard type of jet condenser air pumps but practice recom- 
mends the use of separate air and wet-vacuum pumps for vacua higher 
than 26 inches. 

Since the wet-air pump for surface condenser handles only the con- 
densed steam and air, its theoretical capacity, neglecting clearance, 
may be determined by eliminating V from equation (261) which then 
becomes 

e = „ + „„_2?^x;-i±ip5. (262) 

Vc - Vv ta + 450 

The volume of air entering the condenser varies so much with the 
character of the power-plant equipment and the conditions of operation 
that an}'^ assumed average value of Va may lead to serious error. 

Average steam turbine practice gives 

Q = 20 V for 26-inch vacuum, 
Q = 30 z; for 27-inch vacuum, 
Q = 40 V for 28-inch vacuum, 
Q = 50 V for 29-inch vacuum. 

Average reciprocating engine practice gives 

Q = 85 per cent of above for vacua up to 27 inches. 

308. Tail Pumps. — As previously stated the term ^Hail" pump has 
been applied to pumps which deal with the combined circulating water 
and condensate merely to distinguish between this type and that 
dealing with the entire condenser water supply including the air entrain- 
ment. In practice the terms tail pump and wet-air pump are used 
synonymously. Almost any type of water pump may be used for the 
purpose of withdrawing the combined circulating water and conden- 
sates but the centrifugal pump appears to be the more common in use. 
Quite recently the ''screw-pump" has been developed to a high point 
of efficiency and it is not unlikely that this type may supplant to a 
certain extent the present type of centrifugal pump. A typical tail 
pump installation is shown in Fig. 295. The Leblanc jet condenser, 



658 



STEAM POWER PLANT ENGINEERING 



Fig. 294, and the C. H. Wheeler low-head high-vacuum jet condenser, 
Fig. 296, involve the use of centrifugal tail pumps. The power required 
to drive this style of pump may be calculated from equation (263). In 
this connection the total head pumped against must include the suction 
head due to the vacuum in the condenser. 

309. Dry-air or Dry-vacuum Pumps. — Dry-air or dry- vacuum pumps 
are used in connection with jet or surface condensers where a high 

degree of vacuum is essen- 



r=^i 




Fig. 402. Air Cylinder Construction of Wheeler 
Dry-vacuum Pump. 



tial as in steam turbine 
practice. Such pumps are 
intended to exhaust the sat- 
urated non-condensable va- 
pors only. Air pumps for 
jet condensers must deal 
with much larger volumes 
of air than those for surface 
condensers, other things be- 
ing equal, because of the air 
entrained with the circulat- 
ing water. Dry-air pumps 
(1) the reciprocating piston, 



1st Stage Suction 

^ f 



may be divided into four general groups 

(2) positive rotary displacement, (3) hydro-centrifugal and (4) steam-jet. 

Fig. 402 shows a section 
through the cylinder of a 
W^heeler dry-vacuum pump 
illustrating the single-cylin- 
der, single-stage reciprocat- 
ing] piston group. The ad- 
mission valves A and A are 
mechanically controlled and 
the discharge valves are of 
the usual spring loaded type. 
The rotary admission valves 
are adjusted so that for a 
short instant at dead center 
communication is established 
between both ends of the Fig 
cyUnder so as to reduce the 
air pressure in the clearance 
space down to the suction pressure on the other side of the piston. 

Fig. 403 shows a section through the cylinder of a Worthington 
single-cylinder two-stage dry-vacuum pump and which possesses many 




Connection to Vacuum Trap 

403. Air Cylinder Construction of Worth- 
ington Two-stage Single-cylinder Dry-vacuum 
Pump. 



PUMPS 



659 



advantages over the single-cylinder mechanism. The cycle of operation 
is as follows: With piston moving as indicated air is drawn into the 
head-end of the cylinder until the piston reaches the end of its stroke. 
On the return stroke the air drawn in the head end of the cylinder is 




SECTION BB 



Fig. 404. Leblanc Air Pump. 



transferred (at condenser pressure) through passage D and valve E to 
the crank end of the cjdinder. On the next stroke the air charge is 
compressed through spring loaded valve H to somewhat more than 
atmospheric pressure. 

The Leblanc, Thyssen, Wheeler Turbo-air Pump and the Worthing- 
ton HydrauHc Vacuum Pumps are well-known examples of the hydro- 
centrifugal or hurling-water dry-air pumps. They differ very little 




Fic;. 405. Thyssen Vacuum Pump. 

from each other in principle but vary widely in mechanical construction. 
In these pumps entraining or hurUng water is taken from a circulating 
tank and hurled by centrifugal force in thin sheets or '^ pistons" into a 
diffuser or discharge cone, each sheet or piston carrying with it a layer 
of saturated air drawn in from the condenser. The water is used over 



660 



STEAM POWER PLANT ENGINEERING 




Impeller 



Hurling Water 
Inlet 



and over again since very little heat is abstracted from the air. This 
style of pump is in common use and has superseded the reciprocating 
pump to a great extent. It may be driven by motor or turbine, is very 
compact, and owing to the absence of valves and reciprocating parts 

requires very little attention. The 
power requirements, however, are 
from two to three times that for a 
reciprocating pump. 

Fig. 408 shows an application of 

the Parsons augmenter which is one 

of the earhest applications of a steam 

jet for withdrawing the non-con- 

'compression dcusablc vapors from a condenser. 

Channels 

Referring to the illustration, a pipe 
Fig. 406. Diagrammatic Arrangement is led from the bottom of the main 
of Elements in Wheeler Turbo-air condenser to an auxihary or aug- 
^^™P' menter having about one-twentieth 

of the cooling surface of the main condenser. At the point indicated a 
small jet is provided which acts as an ejector and draws out the air and 
vapor from the condenser and delivers it to the air pump. The water 
seal prevents the air and vapor from returning to the condenser. With 
this arrangement if there is a vacuum of 27| or 28 inches in the con- 
denser there need be only 26 at the air 
pump, which therefore may be of smaller 
size, the jet compressing the air and vapor 
and the augmenter condenser cooUng them 
so that the volume is reduced about one 
half. The steam jet uses about 1| per cent 
of the steam used by the prime mover at 
full load. The net saving on the average 
condenser due to the use of the augmenter 
averages 5 per cent; with light condensers 
the saving is negligible. The kinetic ejector 
is a development of the Parsons vacuum 
augmenter but since it is little used in this 
country no attempt will be made to describe ^^''- ^0^; Worthington Hy- 
it. A notable installation of the kinetic ^^^^ acuum ump. 

ejector is in the Fisk Street Station of the Commonwealth Edison 
Company, Chicago, Illinois. 

Fig. 409 shows a general assembly of the C. H. Wheeler ''Radojet" 
pump which is the latest development of the steam jet for vacuum pur- 
poses and which promises to supersede the hydro-centrifugal pump for 




PUMPS 



661 



general condenser practice. This device consists essentially of a com- 
pound live-steam jet; a primary jet which withdraws the saturated air 
from the condenser and compresses it to four or five inches above 
condenser pressure and a secondary jet which picks up the discharge 
from the primary and forces it out against atmospheric pressure. By 
forcing the discharge into an open feed-water heater the latent heat of 
steam used by the jets may be reclaimed. The primary jet is effected 
by a number of small expanding nozzles discharging into a conical dif- 
fusing chamber. The secondary jet is ra- 
dial in form and discharges into an annular 
volute chamber. There are no moving 
parts and the apparatus is very compact 
and simple. The same degree of vacuum 
may be developed for identical operating 
conditions as with the hydro-centrifugal 
air-pump and at a lower power cost. 





Fig. 408. Parsons Vacuum Augmenter. 



Fig. 409. C. H. Wheeler 
* ' Rado j et ' ' Dry-vacuum 
Pump. 



310. Size of Dry-air Pumps. — The volumetric capacity of a dry- 
air pump for condenser service is based upon experience rather than 
theory because the amount of air in the steam and the air infiltration 
are very uncertain quantities. Since the air to be dealt with is saturated 
with water vapor the pump displacement or its equivalent will be much 
larger than if dry air only were supplied. The volume of mixture 
which must be exhausted for a given weight of dry air for different vacua 
and air-pump suction temperatures is shown in Fig. 410. The curves 
are based on equation (260) and give the volume of mixture containing 
one pound of dry air at various condenser pressures and corresponding 
saturated vapor temperatures. The great reduction in volume effected 
by cooling the air-pump suction is clearly shown. The marked superi- 
ority of counter current over parallel current flow for high vacua is 
chiefly due to the greater reduction in temperature of the air and its 
vapor content. 



662 



STEAM POWER PLANT ENGINEERING 



The following capacities for dry-air pumps appear to conform with 
current practice : 

Q = 20 y to 30 y for vacua under 27 inches. 

Q = 35 ?; to 50 i^ for vacua of 28 inches or over, both referred to a 

30-inch barometer. 
Q = air-pump displacement, cu. ft. per hr. 
V = volume of condensate, cu. ft. per hr. 



1300 






























1400 










^ 




















1200 










\\ 




5 


1 i 














11 nn 








/ 


d 




31 ^ 














^tj 








/ 


2 

'C 


Mil 
^1 














^1000 

o 

T3 
- 000 








/ 


1 
s 






2 










.000 






J 


si 


« 


'11 


si 


i 










X -no 






/ 




s 1 


5 


n 






1 


V 




o 






/ 




•PI 


/ 


1 

1 i 


/ 


« 




3 / 


1 




■g 600 




1/ 




/ 


if 


/ 


1 I 
\ I 
1 \ 




C. 






C 




3 


/' 






/ 


/ 


/ 


1 ; 




5 






u 




4UU 
300 




^ 


y 




/ 




1/ 








/ 













^ 




^^^ 




1 






/ 


' 


1 










— 








1 




X^ 










100 





















1 

1 













30 40 50 60 70 80 , 00 100 110 120 130 140 150 

Temperature of Air Pump Suction, Deuces Fahrenheit 

Fig. 410. Cubic Feet of Saturated Air Containing One Pound of Dry Air for 
Various Vacua and Air Temperatures. 

In a number of recent large condenser installations the air pumps are 
proportioned on a basis of Q = 50 f. 

The curves in Fig. 411, though strictly apphcable to a specific case, 
represent the general characteristics of an ordinary reciprocating vs. 



PUMPS 



663 



a hydro-centrifugal air-pump. Referring to the curves, it will be seen 
that the reciprocating pump is superior to the hydro-centrifugal for 
vacua below the line BB and for vacua above BB the latter is the more 



u 














,^ 














A- 


H: 










-^ 


z' 


-A 












3 


X 


\ 






/ 


' 
















o 

^ 1 

o 






\ 




/ 






















^ 


\ / 














1 

B 


R- 






_\/- 






-P 


















\\ 


















u 

< 






r 


1 \ 


























\ c 
















£ 
5 












Ts 


























1 
















h 








i 




\| 




















< 




\.S 






















1 
































1 



















•Z 3 i 5 6 r 8 9 10 11 12 13 • 
Cubic Feet of Air Removed per Second 

Fkj. 411. Comparative Tests — Reciprocating Air Pumps vs. Leblanc Air Pumps. 

















100^ Vacuum 










"^ 


100< 

























— 









26 






























80 

70 1 
60 1 






























c 27 

S 28 
> 
























































^ 






















^^ 








30 
20 
10 
















^ 














o 29 












L-1 


r 
















1 
30 


— 


==^ 


" 








acu 


im C 


orre 


spon 


ii°j. 


to .0 


^a 


ter 


.__ 































10 



20 



40 



50 



Cubic Feet of Free Air per Minute 

Fig. 412. Test of Wheeler Turbo-air Pump. 

effective. For vacua above AA the hydro-centrifugal pump is in a 
class of its own. With tight condensers in which air leakage is kept to 
a minimum a reciprocating air-pump of the Worthington two-stage 



664 



STEAM POWER PLANT ENGINEERING 



Discharge 



single-cylinder type (Fig. 403) may maintain a higher vacuum than the 
hydro-centrifugal type for the same temperature range. 

311. Centrifugal Pumps. — Centrifugal pumps consist of two essen- 
tial elements, (1) a rotary impeller which draws in the water at its 

center and (2) a stationary casing which 
guides the water thrown from the ends of 
the impeller to the discharge outlet. In- 
crease of peripheral speed increases the 
energy in the impeller. This increase in 
energy may take the form of increase in 
pressure or potential energy, or it may be 
in the form of increase in rate of flow or 
kinetic energy. In general there is an in- 
crease in both kinetic and potential energy. 
The impeller may be of the open type, 
Fig. 414 (B), or closed, Fig. 414 (A). The 
casing may be cyhndrical and concentric 
with the impeller, Fig. 418, or of spiral form. Fig. 413. It may be plain 
or fitted with diffusion vanes and any number of impellers may be 
employed. The shape of the impeller and casing and the number of 
impellers or stages determine the efficiency of the pump and its adapt- 
abihty to certain conditions of service. 

Centrifugal pumps are generally classified as 

1. Volute. 

2. Turbine. 

Fig. 413 gives an end view of a typical single-stage volute pump with 
end plate removed so as to expose the impeller, and Fig. 415 shows a 




Fig. 



Suction 

413. A Typical Centrif- 
ugal Pump. 





Fig. 414. Basic Types of Impellers. 



section through a modern single-stage volute pump with double suction. 
In the volute pump the casing is of spiral design forming a gradually 
increasing water or '^ whirlpool" chamber, A-B, Fig. 409, for the 
purpose of partially converting velocity head to pressure head. The 
older forms of volute pumps were very inefficient, seldom deUvering 



PUMPS 



665 



more than 40 per cent of the energy supplied and usually not adapted 
to lifts greater than 50 feet. The modern pumps give efficiencies as 
high as 80 per cent, and the lift is limited only by the speed of the im- 
peller. As a general rule the volute pump is of single-stage construction 




Fig. 415. Typical Single-stage Double-suction Volute Pump. 

and limited to comparatively low lifts, 120 feet and under, though two- 
stage pumps of this type are on the market designed for heads as high 
as 1000 feet. 





Fig. 416. Direction of Water from the 
Impellers of a Centrifugal Pump 
without DifTusion Vanes. 



Fig. 417. Effect of Diffusion Vanes 
on the Direction of Water. 



In the usual design of volute pumps the stream of water in the casing 
is at cross current with that thrown out from the impeller as shown in 
Fig. 416. The turbine pump is provided with a system of diffusion 
vanes or expanding ducts, disposed between the periphery of the im- 



666 



STEAM POWER PLANT ENGINEERING 



peller and the annular casing, somewhat like the guide vanes in a 
reaction turbine water wheel, so that the fluid emerges tangentially at 
about the velocity in the casing (see Fig. 417). The casing is usually 
concentric with the impeller and of uniform cross section though the 
volute casing is sometimes used in this connection. For high lifts 
these pumps are compounded, thereby reducing the peripheral velocity 
and decreasing the friction losses. Fig. 418 shows a section through a 
three-stage Worthington turbine pump as installed in the testing labora- 
tories of the Armour Institute of Technology and designed to deliver 
200 gallons per minute against a 750-foot head at 2500 r.p.m. 




Fig. 418. Worthington Three-stage Turbine Pump. 

In view of past developments it is probable that the centrifugal pump 
will supplant the piston type of pump for practically all purposes, ex- 
cept perhaps for deep-well service and for very heavy pressures. Cen- 
trifugal pumps are now used for boiler feeding, circulating condensing 
water, hot-well and wet-vacuum purposes and for various applications 
of industrial service. Efficiencies above 70 per cent are not unusual 
and the head against which the pump may operate is limited only by 
the peripheral speed at which the impeller may be safely run. Although 
the equivalent heat efficiency of the high-grade piston pump is superior 
to that of the centrifugal pump, other items, such as low first cost, 
decreased cost of repairs and the like, frequently offset this advantage. 
Some of the advantages of the centrifugal pump as compared with the 
piston type are: 

1. Low first cost, 

2. Compactness, 

3. Absence of valves and pistons, 

4. Low rate of depreciation. 



PUMPS GG7 

5. Uniform pressure and flow of watc^r, 

6. Simplicity of design and ease of operation, 

7. Freedom from shock, 

8. High rotative speed, permitting direct connection to electric 

motors and steam tm^bines, 

9. Abihty to handle dirty water, sewage and the like, 

10. In case of stoppage of delivery, the pressure cannot increase 

beyond the predetermined working pressure, and 

11. Ease of repair. 

Some of the disadvantages are: 

1. Efficiency not as high as the best grade of piston pumps, 

2. Cannot be direct connected to low-speed engines when high lifts 

are desired, and 

3. The rate of flow cannot be efficientl}' regulated for wide ranges in 

duty. 

313. Performance of Centrifugal Pumps. — For best efficiency a cen- 
trifugal pump must be properh' designed for the intended service as 
to curvature of vanes, diameter and speed of impeller, and number of 
stages. Figs. 419 to 421 are based upon experiments Avitli De Laval 
centrifugal pumps. When a practically uniform head is required at 
constant speed with varying water supply as in city water works, hydrau- 
lic elevator systems or boiler feeding, the impeller vanes are designed 
to give the characteristic curve illustrated in Fig. 419 which protects the 
motor from possible overload. See also Fig. 430. 

In dry-dock and other variable head work, in oraer not to overload 
the motor, the power should be practically constant through wide 
variations of head and at the same time the efficiency should not vary 
seriously. A desirable characteristic for such a pump is illustrated in 
Fig. 420. 

In water-supply systems in which the friction of the piping is a large 
part of the total head at full delivery, the characteristic shown in Fig. 
421 is especially useful. Thus, when the system reduces its demand for 
water and the frictional head is consequently considerably reduced, 
the pump would automatically adjust itself to the reduced head without 
change of speed. Figs. 419 and 422 are based upon experiment and show 
the relationship between speed, head, capacity, efficienc}" and power 
consumption of various types of pumps. 

The theory involved in the operation of centrifugal pumps and rules 
for design are beyond the scope of this book and the reader is referred 
to the accompanying bibliography. 



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Fig. 419. Centrifugal Pump Characteristic for Hydraulic Elevator Service, 
Boiler Feeding, etc. 





























































































































































































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Fig. 420. Centrifugal Pump Characteristic for Dry-dock Service. 

























































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Fig. 421. Centrifugal Pump Characteristic for Water Works with Large Friction Head. 

(668) 



PUMPS 



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lea-deq'en centrifugal pump 

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Fig. 423. Performance of Two-stage Lea-Degen Centrifugal Pump. 



670 



STEAM POWER PLANT ENGINEERING 




Fig. 424. Two-lobe Cycloidal 
Pump. 




Fig. 



425. Rotary Pump with 
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Fig. 426. Performance of a Small Rotary Pump. 



PUiNIPS 



671 



313. Rotary Pumps. — Rotary pumps arc often used for circulating 
cooling water in condenser installations, and give about the same effi- 
ciency as centrifugal pumps under similar conditions of operation. 
For moderate pressure and large volumes they offer the advantage of 
low rotative speed, thus permitting direct connection to slow-speed 
steam engines. At high speeds they are noisy, due chiefly to the gear- 
ing. They occupy considerably less space than piston pumps of the 
same capacity, but require more room than the centrifugal type. 























































































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Fig. 427. Test Curves for 8-inch "Screw" Pump, American Well Works. 

Fig. 424 shows a section through a two-lobe cycloidal pump. The 
shafts are connected by wheel gearing, the power being applied to one 
of the shafts. The water is drawn in at / and forced out at 0, the 
displacement per revolution being equal to four times the volume of 
chamber A. There is no rubbing between impellers and casing. In 
this type of pump the pressure is independent of the speed of rotation, 
and the capacity varies almost directly Avith the speed. The slip varies 
from 5 to 20 per cent according to the discharge pressure. 

Fig. 425 shows a section through a rotary pump with movable hut- 
ment. Fig. 426 illustrates the performance of a 45-mm. Siemens- 
Schuckert rotary pump at different speeds and discharge pressures. 
(Zeit. d. Ver. Deut. Ing., June 24, 1905, p. 1040.) Large rotary pumps 
give much higher efficiencies, but the general characteristics are about 
the same. A combined efficiency of pump and engine as high as 84 
per cent has been recorded. (Trans. A.S.M.E., Vol. 24, p. 385.) 



672 



STEAM POWER PLANT ENGINEERING 



Screw pumps may be grouped with the rotary positive-displacement 
class. The Quimby screw pump is one of the best-known examples of 

this type of pump and 
consists essentially of two 
right and left square 
thread screws revolving in 
a double casing. The 
liquid to be pumped is 
drawn in at the outer ends 
of the cyhnder and forced 
toward the center by the 
action of the two pairs of 
intermeshing threads. The 
discharge is from the cen- 
ter of the casing. Power 
is appUed to one of the 
screws and the second is 
driven by means of a pair 
of gears. The screws run 
in close fit with the casing 
but without actual con- 
tact. Quimby pumps 
operate at speeds varying 
from 600 to 1500 r.p.m., 
depending upon the size 
and service for which they 
are intended. Fig. 427 
shows the performance of 
an 8-inch ''screw" pump 
built by the American 
Well Works. 

314. Circulating Pumps. 
— This term is ordinarily 
applied to the pumps 
which supply cooling water 
to surface condensers. The 
three types found in con- 
denser practice are (1) the 
centrifugal, (2) the rotary 
positive-displacement, and 
(3) the reciprocating-piston pump. The centrifugal pump is by far the 
more common in use. For high lifts and in connection with very large 




10,000,000-gallon Circulating Pump. 






PUMPS 



673 



units the high-duty reciprocating j^iston pump has been used 'because 
of its high overall efficiency but such installations are exceptional. 
The rotary pump is occasionally used where the driving unit is a slow- 
speed reciprocating engine. In small and medium-sized installations 
the screw pump has also been used but in the majority of plants the 
centrifugal pump appears to be the best selection. 

The power required b}' the circulating pumps is the largest item of the 
condenser auxiliaries, and therefore every effort should be made to 
reduce the pumping head to 
a minimum. Where it is pos- 
sible to seal the circulating 
water discharge pipe the sys- 
tem operates as a siphon and 
the static head is the difference 
in level of intake and dis- 
charge canals. Where the dis- 
charge head cannot be sealed 
the static head is the difference 
in level of intake water and 
the top pass in the condenser. 
The total head pumped against 
in any case is the sum of the 
static head (suction plus dis- 
charge) and the friction head 
lost in the condenser and piping, 
deliver the circulating water is 



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Typical Centrifugal Pump. 

The brake horsepower necessary to 



Br.hp. = - 



WH 



(263) 



33,000 E 
in which 

W = weight of circulating water, lb. per min., 

H = total head, ft., 

E = mechanical efficiency of the pump. 

The static head of course remains constant, other conditions being 
the same, for all rates of flow, but the friction head increases with the 
square of the quantity pumped. This is illustrated in Fig. 429. 

Example 63. Calculate the power required to drive the circulating 
pump for a surface condenser installation when operating under the 
following conditions: Maximum capacity of main tur})ine 10,000 kw., 
water rate 15 lb, per kw-hr., ratio of cooling water to condensate 60, 
suction head 5 ft., friction heat 20 ft., static discharge head 15 ft., 
pump efficiency 70 per cent. 



674 



STEAM POWER PLANT ENGINEERING 



From equation (263), 

15 X 10,000 X 60 (5 + 20 



Br.hp. = 



15) 



= 261 (approx.). 



60 X 33,000 X 0.7 

If the pump is motor driven allowing an overall motor efficiency of 
85 per cent the pump will require 

261 _ ( 0.023 or 2.3 per cent of the main 

10,000 X 1.34 X 0.85 ~ ( generator output. 



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Fig. 430. Typical Performance Curves of a Rees "Roturbo" Boiler-feed Pump. 

315. Centrifugal Boiler-Feed Pumps. — In power plants having capaci- 
ties over 1000 boiler horsepower direct-acting and power-driven triplex 
boiler-feed pumps have been largely superseded by turbine- or motor- 
driven centrifugal pumps. For plants under 1000 
horsepower the direct-acting pump offers the advan- 
tage of low first cost and^ ease of operation. The 
most economical drive for a centrifugal boiler-feed 
pump is a steam turbine using the exhaust steam 
for heating the feed water, though motor drives are 
sometimes used to advantage in large central sta- 
tions. One great advantage of a steam turbine- 
driven centrifugal pump is that its delivery may be 
throttled down to zero when the pump is operating 
at its normal speed. A further advantage is that it 
delivers a uniform and even supply without pulsa- 
tions or the need of air chambers or relief valves, 
thus avoiding vibration and water hammer. The 
turbine-driven pump is occasionally equipped with a 
water-pressure governor which regulates the speed of the turbine and 
adjusts it automatically to any load. The power required to dehver 
w^,ter to the boiler maj^ be calculated with the aid of equation (263). 
In practice the centrifugal boiler-feed pump requires from less than 




Fig. 430a. The 
Pulsometer. 



PUMPS ()7r) 

one per cent to five per cent of the boiler steam depending upon the 
size, load, type of drive and disposition of the exhaust. 

Example 64. Calculate the power required to drive the centrifugal 
feed pump for a turbine installation when operating under the follow- 
ing conditions: Maxinmm output of main turbine 10,000 kw., water 
rate (including auxiliary steam) 16 lb. per kw-hr., boiler pressure 200 
lb. gauge. 

When specific figures are not available it is customar}^ to assume 
25 per cent of the boiler pressure as the friction head, whence H = 
(200 + 50) 2.6 = 650 ft. (2.6 = ft. of water at boiler temperature cor- 
responding to 1 lb. per sq. in.). Assume a pump efficiency of 65 per cent. 

From equation (263), 

^ ^ 16 X 10,000 X 650 ^ 

P' ~ 60 X 33,000 X 0.65 ~ 

If the pump is turbine driven and the latter used 40 lb. of steam per 
b.hp-hr. the pump will require 

81 X 40 



160,000 



0.02 or 2 per cent of the total weight of steam generated. 



If the pump turbine exhaust is used for feed water heating the pump 
will require only 0.3 per cent of the total steam generated. (See example 
58.) 

316. Condensate or Hot-well Pumps. — The centrifugal pump is now 
quite universally used for pumping the condensate from surface con- 
densers. Condensate pumps must deliver water against the head cor- 
responding to the vacuum, plus the friction head and the static head. 
The pump cannot create a vacuum sufficiently greater than the vacuum 
in the condenser to draw water into the impeller by suction, therefore 
the condensate should be suppHed under a head of three or four feet or 
more. If the head on the suction side is less than this the pump '^cavi- 
tates" or becomes vapor bound and is unable to remove the water. 
Condensate pumps are built in single-stage and two-stage types. These 
pumps are ordinarily operated without automatic control and are per- 
mitted to operate at constant speed. The power required to operate 
the pump may be calculated with the aid of equation (263). 

Example 65. Calculate the power required to drive the condensate 
pump for a turbine installation when operating under the following 
conditions: Maximum output of main turbine 10,000 kw., water rate 
15 lb. per kw-hr., vacuum 28 inches referred to a 30-inch barometer. 

Suction head corresponding to 28 in. of mercury = 31 ft. 

Assume a friction and discharge head of 29 ft. ; efficiency 50 per cent. 

Substituting these values in equation (263), 

^ . 10,000 X 15 X (31 -h 29) ^ . , , 

^^•^P- = "T o X 33,000 X 0.5 = ^-^ ("PP^""-^- 



676 



STEAM POWER PLANT ENGINEERING 



317. Air Lift. — The air lift is a simple arrangement of piping where- 
by water may be raised by means of compressed air. There are no work- 
ing parts, and no valves are employed except to regulate the supply of 
air. Its particular field of application lies in pumping water from a 
number of scattered wells, and on account of the total absence of work- 
ing parts it is peculiarly adapted to handling water containing sand, 
grit and the like. The device consists of a partially submerged water 
pipe and air supply variously arranged as in Fig. 431 (A) to (D). Com- 
pressed air forced into the water pipe at or near the bottom decreases 
the density of the column and the difference in weight between the 




"Water Level 
in Well 



(A) '' (B) (C) (D) 

Fig. 431. Various Arrangements of the ''Air Lift." 



solid column of water B and the air- water column A causes the flow 
The successful operation of this device depends upon the ratio of the 
depth of submersion B to the total head C. 

The quantity of air necessary to operate an air lift may be closely 
approximated from the equation (see Prac. Engr. U. S., April 1, 1912, 
p. 354) ^ 

log-3;^ X C 
in which 
V = cubic feet of free air per gallon, 
*S = actual submergence in feet, 
C = coefficient determined from experiment. 
The actual submergence *S may be determined from the relationship 



^ LSp 
ip 



(265) 



PUMPS 

in which 

L = actual lift in feet (A, Fig. 431), 



67: 



Sp = submergence percentage ( 100 -r, , Fig. 431 ], 
Ip = lift percentage ^00 ^, Fig. 431 V 



The coefficient C may be approximated as follows: 

C = 255 - 0.1 L. (266) 

For the air pressure required for any lift and any percentage of sub- 
mergence it is convenient to divide the actual submergence in feet by 
2 to get the gauge pressure in pounds. This gives enough pressure in 
excess of that due to water head to allow for the pipe friction and other 
losses. 

The efficiency (''water" horsepower divided by "air" horsepower) 

varies from 30 to 50 per cent, increasing as the ratio -^ increases from 

0.55 to 0.85. (Engineer, U. S., Aug. 15, 1904, p. 564.) A number of 
tests gives efficiencies (''water" horsepower divided by i.hp. of steam 
cylinder) varying from 20 to 40 per cent. The horsepower required 
to compress one cubic foot of free air to different pressures per square 
inch, as determined from actual practice, is approximately as given in 
Table 115. 

TABLE 115. 



Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


176 

140 

100 

80 


0.434 
0.376 
0.201 
0.189 


60 
45 
30 


0.159 
0.145 
0.121 



(Engr., Lond., Aug. 14, 1903, p. 174; Dec. 11, 1903, p. 568; Feb. 12, 1904, p. 172.) 

When it becomes necessary to raise water to a height exceeding sa}^ 
175 feet above the level in the well, it is customary to use two or more 
pumps, the total lift being divided between them. 

Air Lift: Power, June 22, 1915, p. 843; Eng. and Contr., Aug. 9, 1916, p. 137; 
Bulletin No. 450, Univ. of Wis.; Prac. Engr., April 1, 1912. 

Pulsometer: Tech. Quar., Sept., 1901; Public Works, Aug. 15, 1904; Engr. U. S., 
July 15, 1904; Experimental Eng., Carpenter, p. 621. 

Turbine Pump Design: Jour. A.S.M.E., Sept., 1915., p. 538. 



678 STEAM POWER PLANT ENGINEERING 

New Centrifugal Pump Duty Record: Iron Age, Apr. 20, 1916, p. 940. 

Centrifugal Boiler Feed Pumps: Power, Oct. 31, 1916, p. 609; Aug. 24, 1915, 
p. 276; Nov. 16, 1915, p. 693; Dec. 29, 1914, p. 934. 

Characteristic Curves of Centrifugal Pump: Jour. W. Soc. Eng., Oct., 1914, p. 776. 

Pumping Units of Various Types for Small Water Supply Systems: Munic. Jour., 
June 22, 1916, 879. 

PROBLEMS. 

1. A direct-acting duplex boiler-feed pump uses 125 lb. steam per i.hp-hr. Initial 
steam pressure 115 lb. absolute, feed-water temperature 180 deg. fahr. What 
per cent of the total steam generated by the boiler is necessary to operate the pump? 

2. A triple-expansion pumping engine delivers 30,310,000 gallons of water in 24 
hours against a head of 61 lb. per sq. in., initial steam pressure 200 lb. abs., developed 
hp. 800, water rate 10.33 lb. per br.hp-hr., steam initially dry. Required the duty 
per 1000 lb. of dry steam and per million, B.t.u. 

3. Determine the cylinder dimensions of a direct-acting single-cylinder feed 
pump suitable for a 5004ip. boiler, maximum overload 100 per cent, boiler pressure 
115 lb. abs., feed-water temperature 70 deg. fahr. 

4. Required the probable i.hp. when operating at maximum capacity. 

5. Which is the more economical in heat consumption as a boiler feeder, an in- 
jector or a motor-driven triplex power pump? Boiler pressure 100 lb. abs., feed water 
supply 60 deg. fahr., injector delivers 16 lb. of water per lb. of steam, overall effi- 
ciency of pump and motor 60 per cent. 

6. Approximate the cylinder dimensions of a wet-air pump for a 750-hp. engine 
using 16 lb. steam per i.hp-hr., initial pressure 150 lb. abs., vacuum 26 in. (barometer 
30 in.), dry steam at admission, initial temperature of injection water 70 deg. fahr. 

7. Required the horsepower necessary to op3rate a centrifugal circulating pump for 
a surface condenser installation using 1000 gallons of water per minute, total head 
pumped against 50 ft., initial temperature of circulating water 70 deg. fahr. 

8. If the pump in Problem 7 is installed in connection with a 1000-hp. engine 
and the ratio of coohng water to condensed steam is 30 to 1, required the per cent of 
main engine power necessary to operate the pump. 

9. If the pump in Prc'^lem 8 is driven by a steam engine using 50 lb. steam per 
hp-hr. and the exhaust is "^ed for heating the feed water, required the per cent of 
main engine heat supply necessary to operate the pump. Main engine initial pres- 
sure 150 lb. abs., vacuum 26 in. (barometer 30 in.), circulating-pump engine initial 
pressure 100 lb. abs., back pressure 16 lb. abs. Assume dry steam at admission in 
both cases. 



CHAPTER XIV 

SEPARATORS, TRAPS, DRAINS 

318. Live-steam Separators. General. — The function of a steam sepa- 
rator is the removal of entrained water from steam. 

Unless a boiler is hberally provided with superheating surface, the 
steam may contain an amount of moisture varying from 0.3 to 5 per 
cent. If the boiler is poorly proportioned or forced far above its rating, 
this percentage may be greatly increased. The quality of the steam 
is still further reduced by condensation in the steam pipe, which may 
vary from 1 to 10 per cent, depending upon the length of pipe and effi- 
ciency of covering. 

One of the effects of moisture in steam is to increase its density and 
reduce its elastic force. It also increases its conductivity, so that 
during the work of expansion more heat is absorbed from the walls of 
the cyHnder and discharged into the atmosphere or into the condenser 
without doing useful w^ork. (Ewing, ''The Steam Engine," p. 151.) 
Although the heat loss from this cause is small, the danger arising from 
the introduction of a considerable amount of water in the cylinder renders 
the removal of the moisture necessary. See par. 193 for influence of 
moisture on steam consumption. 

The essentials of a good separator are high efficiency as a water 
eliminator, ample storage capacity for any sudden influx of water, 
simphcity and durability in construction, and small resistance to the 
current of steam passing through. A good separator may be relied 
upon to remove practically all of the moisture from steam containing 
under ten per cent entrainment and all but two per cent from steam 
containing as much as twenty per cent. (Engineer, U. S., Jan. 15, 
1904.) 

Table 116 gives the results of a series of tests made by Professor R. C. 
Carpenter in 1891 of six steam separators. (Power, July, 1891, p. 9.) 
Conclusions from these tests were: 

1. That no relation existed between the volume of the several sepa- 
rators and their efficiency. 

2. No marked decrease in pressure was shown by any of the separa- 
tors, the most being 1.7 pounds by separator E. 

3. Although changed direction, reduced velocity, and perhaps cen- 

679 



680 



STEAM POWER PLANT ENGINEERING 



trifugal force are necessary for good separation, still some means must 
be provided to lead the water out of the current of the steam. 

A series of tests made at Armour Institute of Technology in 1905 on 
a number of separators showed that the efficiency of separation decreased 
as the velocity of the steam increased."^ At the low velocity of 500 feet 
per minute all separators were equally efficient, at a velocity of 5000 
feet per minute several had little effect on eliminating the moisture 
present, and at a velocity of 8000 feet per minute only one gave efficient 
results. 



TABLE 116. 

TESTS OF STEAM SEPARATORS. 
(R. C. Carpenter.) 



Make of 
Separator. 



B 
A 
D 

C. 
E 
F 



Test with Steam of about 10 


Per Cent of Moisture. 


Quality of 
Steam 


Quality of 
Steam 


Efficiency. 


Before. 


After. 




Per Cent. 


Per Cent. 


1 
Per Cent. 


87.0 


98.8 


90.8 


90.1 


98.0 


80.0 


89.6 


95.8 


59.6 


90.6 


93.7 


33.0 


88.4 


90.2 


15.5 


88.9 


92.1 


28.8 



Tests with Varying Moisture. 



Quality of 
Steam Before. 



Per Cent. 

66.1-97.5 

51.9-98 

72.2-96.1 

67.1-96.8 

68.6-98.1 

70.4-97.7 



Quality of 
Steam After. 



Per Cent. 

97.8-99 

97.9-99.1 

95.5-98.2 

93.7-98.4 

79.3-98.5 

84.1-97.9 



Average 
Efficiency 



Per Cent. 
87.6 
76.4 
71.7 
63.4 
36.9 
28.4 



319. Classification of Separators. — Separators are based on one or 
more of the following principles of action: 

1. Reverse current. The direction of the flow is abruptly changed, 
usually through 180 degrees. This causes the water in the steam, on 
account of its greater specific gravity, to be thrown into a receiving 
vessel, while the steam passes on in a reverse direction. 

2. Centrifugal force. A rotary motion is imparted to the steam 
whereby entrained water particles are eliminated by centrifugal force. 

3. Baffle plates. The flow is interrupted by corrugated or fluted plates 
to the surfaces of which the water particles adhere and from which 
they fall by gravity te the well below. 

4. Mesh. The separation is brought about by mechanical filtration 
through screens or meshes. 

The following outline shows the classification of typical separators, 
in accordance with the above principles: 



* See Power, May 11, 1909, p. 834. 



SEPARATORS, TRAPS, DRAINS 



081 



Live-steam separators 



Exhaust-steam separators , 



Reverse current \ g^PP^^^„ 

L otratton. 

( Keystone. 
Centrifui;al < Mosher, 

( Robertson. 

( Bundy. 
Baffle plate < Austin. 

( Detroit. 
Mesh ^ Direct. 

^^^^^^ \ Potter. 

( Jacketed baffle Baum. 

I Absorption Loew. 

320. Types of Separators. Reverse-current Steam Separators. — Fig. 432 
shows a section through a Hoppes steam separator and illustrates the 
principle of reverse-current separation. Steam may flow through in 
either direction. Both the inlet and outlet ports are surrounded by 





Fig. 432. Hoppes Steam Separator. 



Fig. 433. Stratton Steam Separator. 



gutters C, C, partly filled with water, which intercept the moisture follow- 
ing the surface of the pipe, while the downward plunge of the steam 
throws the entrained water to the bottom of the separator. The con- 
densation is carried from the troughs by pipe P to the well below, from 
which it is trapped at D in the usual way. The velocity of the steam in 
passing through this separator is greatly reduced to prevent the steam 
from taking up the water in the bottom of the well. This is brought 
about by increasing the area of the passage through the separator. 

Fig. 433 gives a sectional view of a Stratton separator, which, though 
primarily of the reverse-current type, embodies also the principle of 
centrifugal force. The separator consists of a vertical cast-iron cylinder 
with an internal central pipe C extending from the top downward for 



682 



STEAM POWER PLANT ENGINEERING 



about half the height of the apparatus, leaving an annular space be- 
tween the two. The current of team on entering is deflected by a 
curved partition and thrown tangentially to the annular space at the 
side, near the top of the apparatus. It is thus whirled around with all 
the velocity of influx, producing the centrifugal action which throws 
the particles of water against the outer cylinder. These adhere to the 
surface, so that the water runs down continuously in a thin sheet around 
the outer shell into the receptacle below. The steam, following in a 
spiral course to the bottom of the internal pipe, abruptly enters it, and 
passes upward and out of the separator without having once crossed 
the stream of separated water. The rapid rotation of the current of 
steam imparts a whirling motion to the separated water which tends to 





Fig. 434. Keystone Steam Separator. 



Fig. 435. Bundy Steam Separator. 



interfere with its proper discharge from the apparatus. The separator 
has therefore been provided with wrings or ribs E projecting at an acute 
angle to the course of the current, which have the effect of breaking up 
this whirling motion and allowing the water to settle quietly at the 
bottom, whence it passes off through the drain pipe D. 

Centrifugal Steam Separators. Fig. 434 shows a section through a 
Keystone or Simpson's centrifugal separator. The separator consists of 
a cast-iron cylinder with vertical pipe C extending downward about two- 
thirds of the whole length; this pipe has a thread or screw wound 
spirally around it, the space between the threads being somewhat greater 
than the area of the steam pipe. The steam passing around the spiral 
course causes the water to be thrown against the outer walls by centrif- 
ugal force, while the dry steam passes through the small holes in the 



SEPARATORS, TRAPS, DRAINS 



683 



central pipe. The water passes down the outer walls, where its motion 
is arrested by obstructing ribs E, and is thence carried away by a drip 
pipe D to a suitable drain. 

Baffle-plate Steam Separators. — Fig. 435 gives an interior view of a 
Bundy separator and illustrates the application of baffle plates for live- 
steam separation. This separator consists of a rectangular cast-iron 
casing with a cylindrical receiver beneath it. Directly across the steam 
passage are baffle plates corrugated for the reception of entrained 
water. The plates consist of vertical castings, each containing a main 
artery or channel which leads directly to the receiver. The fronts of 
the plates are flat, with a series of recesses sloping inwards and down- 
wards, terminating in an opening of capillary size leading to the main 
artery. The plates are staggered, so that the 
steam must impinge against all of them in its 
passage. The particles of water adhere to the 
plates, collect, and fall by gravity into the re- 
ceiver. The flanges at the bottom constrict the 
opening of the reservoir so as to prevent the 
steam from picking up an}^ portion of the water. 
Fig. 436 show^s a section through an Austin 
separator and illustrates another class embodjdng 
the fluted baffle-plate principle. The steam in 
passing through the chamber impinges against 
the fluted baffle plate B, The moisture adheres 
to the surfaces, collects and trickles along the 
corrugations to the bottom of the well. These 
corrugations are formed in such a manner that '' ^^^' ^^^tin Steam 
the steam cannot come in contact with the water epara or. 

particles after they have been once eliminated. A perforated diaphragm 
D prevents the water in the well from coming in contact with the 
steam. The current of steam is also reversed, thus giving additional 
separating properties to the apparatus. 

Mesh Separators. — Fig. 437 shows a section through a "direct" 
separator, illustrating the principle of mesh separation. These separa- 
tors are made with steel bodies and cast-iron heads and bases, in all sizes 
up to six inches inclusive, the larger sizes being constructed of cast iron 
or boiler plate. The cone C, perforated lining E, and diaphragm S are 
made of cold-rolled copper; the cone is a substantial gray-iron casting, 
resting on three cast-iron supports hooked over the top of inner pipe as 
indicated. The method of operation is as follows: The accumulated 
moisture around the walls of the steam pipe is caught by the upper edge 
of cone C and carried down back of lining E to the water chamber. The 




684 



STEAM POWER PLANT ENGINEERING 




current of steam entering the separator impinges upon the conical surface, 
which is composed of sohd plate covered with sieve S, through which 
water may freely pass but from which it cannot readily escape. Passing 
through the sieve and depositing on the solid surface of the cone 0, this 
water is carried by conductors P to the water 
chamber. Perforated lining E permits the mois- 
ture content of the steam to pass through the 
opening to the water below and prevents it from 
coming in contact again with the current of 
steam. A trough is provided at the lower edge 
of the inverted cup which leads all the water 
that may adhere to it to the water chamber. The 
steam flows through the passages indicated by 
arrows and is subjected to a whipsnapping action 
which tends to throw off any remaining mois- 
ture. The perforated plate D prevents the steam 
from picking water out of the water chamber. 

331. Location of Separators. — Live-steam sepa- 
rators may be located 

1. Inside the boiler, 

2. Between boiler and engine, 

3. At the steam chest. 

Where the steam pipe is very short, and particularly in marine and 
locomotive work where the tossing of the boiler induces excessive 
priming, the separator may be placed inside the boiler and its function 
becomes that of a dry pipe. In this location it prevents the water due 
to foaming and priming from passing to the engine, and reduces con- 
densation in the pipe by supplying dry steam. The "Potter mesh'^ 
and the "De Rycke centrifugal" are types of separators designed for 
this service. 

The arrangement of separator between engine and boiler, other than 
at the throttle or inside the boiler, is sometimes necessary for economy 
of space. Where possible, however, the separator should be placed 
close to the steam chest. 

Current practice recommends that a receiver separator, which is an 
ordinary separator with a volume of two to four times that of the 
high-pressure cylinder, be placed close to the engine if the load is inter- 
mittent or sharply fluctuating. This forms a cushion for absorbing 
the force of the blows caused by cut-off, delivers steam at a practically 
uniform pressure, and reduces the vibration of the piping to a minimum. 
It also provides a reservoir for sudden demands made by the engine. 
Smaller pipes and higher velocities may be used with this arrangement. 



Fig. 437. ''Direct 
Steam Separator. 



SEPARATORS, TRAPS, DRAINS 



685 



332. Exhaust-steam Separators and Oil Eliminators. — The function 
of an exhaust-steam separator is the removal of cyhnder oil from the 
steam exhausted by engines and pumps. In plants where exhaust 
steam is used for heating it is quite essential to remove the oil from 
the steam before it enters the heating system, for the oil not only re- 
duces the efficiency of the radiators by coating them with an excellent 
non-conducting film but is an element of danger to the boiler itself. 
In condensing plants the separator will prevent the oil from fouling 
the condenser tubes and those of the vacuum heater if one is installed; 
this is an important factor, since the oil or grease lowers the efficiency 
of the heat transmission. 

In a general sense a live-steam separator is also an oil eliminator, and 
all the separators previously described perform this function to a cer- 
tain extent, since the underlying principles governing the elimination 
of oil from exhaust steam are similar to those employed in removing 
water from steam. Most of the separators described above are also 
designed in lighter form, as oil eliminators, but by far the greater 
number are based on the fluted baffle-plate prin- 
ciple, of which the Hine, Bundy, Cochrane, 
Utility, Peerless, and Keily are well-known ex- 
amples. This type of oil separator will elimi- j^^ 
nate a considerable portion of the oil in the 
steam, provided the baffle plates or corrugated 
surfaces are frequently cleaned. 

It is a well-established fact that oil can be 
more effectually removed from wet than from 
dry steam, and some makers, notably the Austin 
Separator Company, inject a cold-water spray 
into the separator chamber. A similar result is 
brought about in the Baum separator, Fig. 438, 
in which the corrugated baffle plate is hollow 
and cold water is forced through the chamber 
thus formed. Referring to Fig. 438: The di- 
verged baffle plate forms the wall of a chamber 
in which cold water is continually circulated. 
This circulation causes moisture to appear on Dram 

the baffle-plate surface. The particles of oil. Fig. 438. Baum Oil 
coming in contact with this moist surface as the Separator, 

steam current is diverged, adhere to it and fall by gravity into the 
well below, where they are completely isolated from the purified steam. 
A large portion of the oil and water, however, does not enter the sepa- 
rator at ah but is caught by the inside ledge near the junction of the 



Water Outlet 




686 



STEAM POWER PLANT ENGINEERING 



exhaust pipe and the separator. The oil and condensation which are 
carried along the bottom of the pipe come in contact with this ledge 
and are carried directly to the outlet pipe. 

A very successful method of removing oil from steam is to project the 
steam on to the surface of a body of water. The water may be hot or 
cold and will hold the oil if it once reaches the surface. It is essential, 
however, to reduce the velocity of the steam as it passes on its way to 
the outlet. Baldwin's grease separator is based upon this principle. 
(Baldwin on Heating, p. 234.) 

The most efficient method of removing oil is by combined filtration 

and absorption. (Engineering News, May 22, 1902, p. 406.) A 

■ large chamber filled with coke, brick, broken 

T [ T tile, or other absorption material is placed in 

y^ ir!~ — N. series with the exhaust pipe. The steam 

' /\ \ passing through this chamber is entirely 

freed from oil and moisture, provided the 
absorbing material is sufficient in quantity 
and is replenished as soon as it becomes 
saturated with oil. The annoyance attend- 
ing the removal and replenishing of the ab- 
sorbing material at frequent intervals and 
the great size of the apparatus are serious 
drawbacks. An example of this system of 
purification in which many of the objection- 
able features are reduced to a minimum is 
the Loew grease and oil extractor. Fig. 439. 
The exhaust steam enters the chamber at 
the top, strikes a large deflecting plate 
shaped like an inverted V, and permits part 
of the condensation and oil to be drawn off 
by the drain pipe. The steam then rises 
and is deflected, as indicated, against a series of shelves filled with 
fibrous material covered with coarse wire screens. The grease is re- 
moved from each shelf by suitable drains. This apparatus is sectional 
and any number of sections may be added without affecting the rest. 
In a non-condensing plant where the exhaust steam is used for heating 
purposes the oil separator is ordinarily placed in the main exhaust 
pipe just before it enters the heating system. Where severoi branches 
enter one main it is not customary to place a separator in e!?.ch branch, 
one large separator located as above being sufficient. 

In condensing plants oil separators are seldom installed except where 
surface condensers are used, in which case the separator may be placed 




439. Loew Grease 
Extractor. 



SEPARATORS, TRAPS, DRAINS 



687 



anywhere between the engine and condenser. In case a vacuum heater 
is used the separator may be placed on either side of the heater, de- 
pending upon the type of separator. If the separator is of the "jacket- 
cooling" or ''spray" type, it may be placed between the engine and the 
vacuum heater; if, however, it is of the ''baffle-plate" type, the oil will 
be more efficiently removed if the separator is placed between the heater 
and condenser so that it will get the benefit of the moisture formed 
in the heater. In the latter location, however, the separator will not 
prevent the oil from fouling the heater tubes. 

Where a jet condenser is used and water is taken from the hot well, 
the hot well itself acts as an oil separator. (Trans. A.S.M.E., 24- 
1144.) 

All separators, steam and oil, should be provided with gauge glasses 
and should be thoroughly drained and the drainage should be 
automatic. 

333. Exhaust Heads. — The function of the exhaust head is the elimi- 
nation of oil and water from steam exhaust before permitting it to be 
discharged into the atmosphere. Unless removed, the water and oil 
rot the roofs and walls in summer and pollute 
the atmosphere surrounding the plant. The ex- 
haust head also acts as a muffler, reducing the 
noise of the escaping steam. Exhaust heads are 
built on the same principle as steam and oil 
separators and most separator builders manufac- 
ture them. Fig. 440 shows a section through a 
typical exhaust head. 

The condensation is ordinarily drained to 
waste, though with proper purification it may be 
returned to the boiler. With an efficient oil sepa- 
rator in the exhaust line the condensation in the 
exhaust head may be returned directly to the 
boiler without further purification. 

Live-steam separators are proportioned so that 
it is only necessary, in the average installa- 
tion, to specify the size of pipe, the type of engine, the steam pres- 
sure, and the style, whether horizontal or vertical. Gauge glasses, 
gauge cocks, and companion flanges are usually provided by the 
maker. In some cases the capacity of the reservoir is also specified. 
In specifying oil extractors the following additional data are neces- 
sary for an inteUigent choice: the number of engines and pumps 
exhausting into the line, the location of the separator, the steam 
pressure, velocity, and the quality and quantity of cylinder oil used. 




Fig. 440. A Typical 
Exhaust Head. 



688 STEAM POWER PLANT ENGINEERING 

A guarantee of efficiency and of material and workmanship is often 
demanded. 

Oil Separation from Water of Condensation: Jour. A.S.M.E., June, 1915, p. 345. 
Electrostatic Separation of Oil and Water: Met. and Chem, Engr., Mar. 15, 1916, 
p. 343. 

334. Drips. — No matter how thoroughly a steam pipe or reservoir 
may be covered with insulating material considerable condensation 
takes place. With the best covering this loss approximates one sixth 
of a pound of steam per square foot of pipe surface per hour for steam 
pressures of one hundred pounds, and runs as high as one pound of 
steam for bare pipes. See Fig. 467 for results of experiments on the 
loss of heat from bare pipes, and Fig. 468 for data on the efficiency 
of pipe coverings. In addition to this water of condensation, from ^ 
to 2 per cent of moisture is carried over by the steam from the boiler. 
This water, unless thoroughly removed, is a constant source of danger 
to the engines and causes water hammer and leaky joints in the piping. 

A joint on a steam pipe may safely withstand a steam pressure of 
100 pounds without leaking and still leak badly under a water pressure 
of half that amount. This is due to the fact that the steam with its 
high temperature causes the pipe to expand, thus insuring a tight 
joint, while the entrained water (which cools as it collects) causes the 
pipe to contract and allows a leak. 

The entrained water and water of condensation are usually spoken 
of as ''drips." Drips may be divided into two classes, low pressure 
and high pressure. 

325. Low-pressure Drips. — Low-pressure drips include the steam 
condensed in heating systems, exhaust steam feed heaters of the close 
type, exhaust steam piping, receiver barrels, steam chests, and exhaust 
heads. As these drips are impregnated with oil and are useless for 
boiler feed without purification, they are usually discharged to waste. 
Most city ordinances require the drips to be cooled to 100 deg. fahr. 
before being discharged into the sewer. In this case they must be 
first discharged into a tank and perixiitted to cool. This tank must be 
vented to the atmosphere to prevent back pressure. Fig. 441 shows 
an installation in which the heat abstracted from the drips, etc., is 
used to heat the feed water. The drips from the throttle valve and 
steam chest in a non-condensing plant are ordinarily discharged into 
the exhaust pipe as shown in Fig. 442. In a condensing plant the 
throttle drips are piped to a trap or to the free exhaust pipe. The re- 
turns from a steam-heating system are sometimes classified as low- 
pressure drips. They are invariably returned to the boiler. 

In small plants all the low-pressure drips may be connected to one 



SEPARATORS, TRAPS, DRAINS 



G89 



large pipe and this pipe in turn to a single trap, provided there is but 
little difference in pressure in the various drip pipes. In case of dif- 
ferent pressures separate leads should be run to waste or traps. 




DraJS^ 




Fig. 441. Closed Heater Installation for Abstracting Heat from Oily Drips. 

The drips from the receiver and vacuum heater barrels in a con- 
densing plant are oftentimes under less than atmospheric pressure, 
and sometimes the pressure varies from 
a slight vacuum to 10 or 20 pounds 
gauge, and consequently cannot be dis- 
posed of as described above. If possible, 
the heaters and receivers should be 
placed so as to drain into the condenser 
(see Fig. 455) . Should this arrangement 
prove impracticable, the barrels may be 
drained by a trap especially arranged as 
shown in Fig. 456. 

336. Size of Pipe for Low-pressure Drips. 
— In the average exhaust-steam feed- 
water heater one pound of steam in con- 
densing gives up approximately 1000 
heat units. This will heat about 6 
pounds of water from 60 to 200 deg. 
fahr. Hence the area of the drip which carries the water of condensa- 
tion from the closed heater need be but one fifth that of the feed pipe. 




Fig. 442. Simple Method of 
Draining Drips. 



690 



STEAM POWER PLANT ENGINEERING 



In no case, however, should a pipe smaller than one half inch in diam- 
eter be used. Should the same pipe be used for both exhaust head 
and heater drips, an area of one fourth area of feed pipe would prove of 
ample capacity. In practice it is customary to use the size of pipe 
conforming with the outlet furnished by the manufacturer of the ap- 
paratus, and only when several pieces of apparatus are connected to 
one main are calculations made for the size of this main. 

The drip pipe from the throttle valve is ordinarily one half inch in 
diameter irrespective of the size of steam pipe; this is also true of the 
steam-chest drip. 

337. High-pressure Drips. — High-pressure drips consist of those 
which are condensed under boiler pressure and include the steam con- 
densed in steam pipes, cylinder jackets of engines, reheating coils of 
receivers, and separators. Being free from oil and containing consid- 
erable heat, they are usually returned to the boiler. Drips may be 
returned to the boiler automatically by means of 

1. Steam traps, 

2. Holly steam loop, 

3. Pumps. 

338. Classification of Steam Traps. — Steam traps may be divided 
into two classes, depending on their use, — return and non-return. 
Both of these two classes may be subdivided into five types according 
to the principle of operation, viz.: 



I. Float. 
11. Bucket. 



III. Bowl. 

IV. Expansion. 



V. Differential. 



CLASSIFICATION OF A FEW WELL-KNOWN STEAM TRAPS. 

i^'-' l^o^o^r'- 

B-ket |A-e.^. 

D-P iMead. 



Steam Traps . 



Expansion 



Differential 



(Metal ISIU^.^- 

■ I Volatile-Fluid | g-h^ 

( Flinn. 
I Siphon. 



Return Traps. 

Traps which receive the condensed steam and return it to a boiler 
having considerably higher pressure than that acting on the returns 



SEPARATORS, TRAPS, DRAINS 



691 



are known as return traps. They are made in a great variety of styles. 
The general principle of operation is shown in Fig. 452 and described in 
paragraph 330. 

Non-return Traps. 

Non-return traps, as the name implies, are used where the water of 
condensation is not returned to the boiler but is discharged into any 
receptacle having less than boiler pressure. 

329. Types of Traps. Float Traps. — Fig. 443 shows a section through 
a McDaniel improved trap, illustrating the principles of the float type. 
A hollow sphere C of seamless copper pivoted at E rises and falls with 
the change of water level in the vessel. The discharge valve M is 
operated by the float. When the trap is empty the float is in its lowest 




Fig. 443. McDaniel Float Trap. 

position and the discharge valve is closed. Water of condensation flows 
into the trap by gravity through opening Z) to a certain depth, when 
the float opens the discharge valve and the steam pressure acting on the 
surface of the water forces it through outlet S to tank or atmosphere. 
After the water is discharged the float closes the valve and permits the 
condensation to collect again. A gauge glass indicates the height of 
water in the chamber. 

Unless float traps are well made and proportioned there is a danger 
of considerable steam leakage through the discharge valve, due to 
unequal expansion of valve and seat and the sticking of moving parts. 
The discharge from a float trap is usually continuous, since the height 
of the float, and consequently the area of the outlet, is proportional to 
the amount of water present. When the trap is working lightly, this 
adjustment is apt to throttle the area and create such a high velocity 
of discharge as to cause a rapid wear of valve and seat. This defect is 



692 



STEAM POWER PLANT ENGINEERING 



more or less evident in all steam traps discharging continuously. For 
this reason aH wearing parts should be accessible and readily replace- 
able. 

Bucket Traps. — Fig. 444 shows a section through an ''Improved 
Acme" steam trap. The water of condensation enters the cast-iron 
vessel at A, filhng the space D between the bucket E and the walls of 
the trap. This causes the bucket to float and forces valve V against 
its seat (valve V and its stem being fastened to the bucket as indicated). 
When the water rises above the edges of the bucket it flows into it and 
causes it to sink, thereby withdrawing valve V from its seat. This 




Fig. 444. Acme Bucket Trap. 

permits the steam pressure acting on the surface of the water in the 
bucket to force the water through the annular space H to discharge 
opening G. When the bucket is emptied it rises and closes valve V 
and another cycle begins. By closing valve R the trap is by-passed 
and the condensation blows directly through passage C to discharge G. 
The discharge from this type of trap is intermittent. 

Dump or Bowl Traps. — Fig. 445 shows sections through a Bundy 
bowl trap of the ''return" type. The water enters the bowl through 
trunnion D and rises until its weight overbalances counterweight E 
and the bowl sinks to the bottom. As the bowl sinks, arm G, which is 
a part of the bowl, rises and engages the nuts N on valve stem H and 
opens valve /, thus admitting live steam pressure on to the surface of 



SEPARATORS, TRAPS, DRAINS 



693 



the water. The trap then discharges hke all others. After the water 
is discharged weight E sinks and raises bowl A, which in turn closes 
valve /, and the cycle begins again. Bowl traps are necessarily in- 
termittent in their discharge. 




Fig. 445. A Typical Tilting Trap. 

Fig. 456 shows the appUcation of a bowl trap to a receiver where the 

drips are under a vacuum, and Fig. 457 a similar application to an 

engine receiver where the pressure varies from less than atmospheric 

pressure to a pressure of 40 or 50 pounds. 

. Expansion Traps. — Expansion traps may be divided into two groups : 

(1) Those in which the discharge valve is operated by the relative 
expansion of metals and 

(2) Those in which the action of a volatile fluid is utiHzed. 
Expansion traps will never freeze, as they are open when cold and all 

the water drains out before the freezing temperature is reached. 



Inlet" 




Fig. 446. A Typical Expansion Trap. 

Since traps of this type have httle capacity for holding water, 5 to 
10 feet of pipe should be provided between the trap and the pipe to be 
drained in order that the condensation may collect and cool. 

Fig. 446 shows the general appearance of a Columbia expansion trap 
in which the valve is operated by the expansion of metaUic tubes. 



694 



STEAM POWER PLANT ENGINEERING 




Fig. 447. Geipel Expansion Trap. 



Water gravitates to the trap through opening marked '4nlet," passes 
through brass pipe 0, then downward to the main body of the valves 
and back to outlet valve C. Below pipe and parallel to it is an iron 
rod S, at the end of which is the support or fulcrum of lever R. The 
lower end of this lever is connected to the stem of the valve C, so that 
any movement of the lever is communicated to it. When the trap is 

cold, valve C is open and 
all water of condensation 
passes out. The moment 
steam enters the pipe it 
expands. The amount of 
expansion is multiplied sev- 
eral times by the action of 
the lever R, so that the 
movement of the valve is much greater than the expansion of the pipe 0. 
The compensating spring D prevents the brass tube from damaging 
itself by excessive expansion. Lever A permits the trap to be blown 
through by hand. 

Fig. 447 shows a section through a Geipel trap in which the valve 
is operated directly by the expansion of two metallic tubes and the 
movement is not multiphed by levers as with the Columbia. The 
lower or brass pipe constitutes the inlet and is connected to the vessel 
to be drained; the upper or iron pipe is the outlet for discharge. The 
two pipes form the sides of an isosceles triangle, the base F of which is 
rigid, while the apex A is free to move in a direction at right angles to 
the linear expansion of the tubes. When cold, the brass pipe is con- 
tracted and the apex, in which the valve seat is placed, is moved down 
so that the valve is open and the water 
is discharged. As soon as steam enters 
the brass pipe the latter expands and 
forces the valve seat against the valve. 
The trap may be adjusted for any pres- 
sure by means of the lock nuts E. 
When it is desired to blow through, the 
valve may be operated by hand by press- 
ing the lever. 

Fig. 448 shows a section through a Dunham trap. It operates upon 
the expansion principle, utilizing a fluid of a volatile character as its 
motive force. The corrugated bronze disk B is filled with a volatile 
fluid, and expands and contracts according to the pressure exerted by 
the fluid. The water enters at the top, surrounds disk B and passes 
through valve opening D to discharge outlet at E. As soon as steam 




Fig. 448. 



Dunham Expansion 
Trap. 



SEPARATORS, TRAPS, DRAINS 



695 



strikes the disk B the volatile fluid flashes into a vapor and causes the 
disk to expand. This expansion forces valve D against its seat and 
the discharge ceases. The valve will remain closed until the condensa- 
tion collects and cools the disk B, which then contracts, opens the valve, 
and condensation enters as before. The adjustment, however, is such 
that the discharge may be made continuous instead of intermittent. 

The Dunham trap is claimed to be the smallest trap of its capacity 
on the market. The 1-inch size, having a capacity for draining 10,000 
hneal feet of 1-inch pipe under 60 pounds pressure, weighs but 5 pounds 
and may be connected to the pipe Hne as if it were a globe valve. 

Fig. 449 shows an internal view of a Heintz steam trap. This works 
on the principle of the volatile-fluid expansion trap but in a different 
manner from any of those described above. The requisite movement 




Fig. 449. Heintz Expansion Trap. 

is obtained by the elongation and contraction of the extremities of a 
bent metallic tube T filled with a highly volatile fluid. This tube is 
inclosed in a cast-iron box and presses against the point of regulating 
screw P. The other extremity of the tube carries the valve and is 
free to move under the action of the variations of temperature. Spring 
*S has no connection with the action of the trap. It is used as a simple 
means of holding one end of the expansion tube on its pivot. The trap 
operates as follows : Water enters at /, surrounds the tube T and passes 
through the valve to the discharge outlet 0. As soon as steam enters 
the chamber the volatile fluid in the tube flashes into a vapor and the 
pressure thus created tends to straighten out the tube; this forces the 
valve against its seat and the discharge ceases. As the trap cools, 
the tube returns to its normal position and the discharge valve is opened, 
thus permitting the condensation to drain out. The adjustment per- 
mits of continuous or intermittent discharge and of variable pressures. 
Differential Traps. — Fig. 450 shows a cross section through a Flinn 
differential trap. The column of water X acting on diaphragm D closes 
valve V. The water entering pipe E and the action of the spring 
equalize column X and open the valve. Describing the action in further 



696 



STEAM POWER PIANT ENGINEERING 







Outlet 



detail, the water of condensation enters at A, fills lower chamber Y, 

pipe X, and receiving chamber C up to the level of the top of pipe E. 

This column of water acting on the under side of the diaphragm D 

forces the valve to its seat against the counter pressure of the spring S. 
Any additional water that enters the trap o\er- 
flows through pipe E, filling chamber F and pipe 
E to a point about midway of its height, where 
the effect of the column of water in pipe X is 
balanced. The pressure on each side of the dia- 
phragm is then equal, the short column in pipe E, 
aided by the spring, balancing the pressure of the 
longer column in pipe X. Any further increase in 
the height of the water in pipe E causes a depres- 
sion of the valve V, which allows water to escape 
until the column has fallen to a level a little below 
the middle of pipe E, when this valve closes again. 
This action is repeated at inter- 
vals according to the quantity 
of water entering the trap. So 
long as the water keeps coming 

Fig. 450. Flinn Dif- in sufficiently large quantities 

ferential Trap. ^^le valve remains wide open. 

Fig. 451 gives a general view of a siphon trap 

which is much used in draining low-pressure sys- 
tems, as, for example, the separator in an exhaust 

steam heating system. It consists essentially of two 

legs A and B, which may be close together or any 

distance apart but the lengths of which must be 

sufficiently great to prevent pressure acting through 

pipe I from forcing the water out of 5. C is a vent 

pipe extending to the air to prevent siphoning; is 

the discharge for the condensed steam. In ordinary 

operation the leg B is filled with water which is 

constantly overflowing, and A with steam and water, 

the total pressure in both legs being equal. The 

siphon trap is apphcable for low pressure only, as it 

requires approximately 2.3 feet of vertical space E 

for each pound per square inch pressure in the pipe. The maximum 

allowable head is represented by vertical distance N. 

330. Location of Traps. — Wherever possible a trap should be 

located so that the condensation will flow into it by gravity. This 

will insure positive drainage. Sometimes, however, the coils, cylinders, 



m 



#^ 



Q Ij Dxa'in 

Fig. 451. Simple 
Siphon Trap. 



SEPARATORS, TRAPS, DRAINS 



097 



or pipes to be drained 'are located in a pit or trench or lie on a base- 
ment floor where it is impossible to set the trap so as to receive the 
drains by gravity without placing it in an inaccessible position. With 
very low pressures this is often unavoidable, but with pressures of 
five pounds or more the trap may be placed above the point to be 
drained. If a trap is set in an exposed place a drain should be pro- 
vided at the lowest point to free the pipe of water when steam is shut 
off. A dirt catcher or strainer should be placed in the pipe leading 
to the trap to prevent scale, etc., from reaching the valve. All pockets 
and dead ends should be drained, and no condensation should be al- 
lowed to accumulate. High- and low-pressure drips should be kept 
separate. All tanks should have gauge glasses. 



steam Snpply 



Equalizing Val 




Fig. 452. Return Traj). 

Fig. 452 shows the application of a float trap for automatically re- 
turning water to the boiler. For this purpose the trap must be placed 
three feet or more above the water line in the boiler, so that the water 
may gravitate to the latter. Water is forced into the trap from the re- 
turns through pipe A until it reaches a level where the float opens the 
equalizing valve V and permits steam from the boiler to enter the trap, 
thus equalizing the pressures. The water then flows into the boiler 
by gravity through check valve D. At the end of discharge the float 
closes the equalizing valve and another cycle begins. Check valve C 
prevents the water from being forced back to the return pipe. If the 
pressure in the return pipe A is not sufficient to force the water into 
the trap, a pump or another trap may be used to effect this result. 



698 



STEAM POWER PLANT ENGINEERING 



Practically any high-pressure trap may be converted into a return 
trap by the proper installation and an ''equalizing" valve. 

Figs. 453 and 454 show different applications of steam traps to the 
receiver coils and jackets of triple-expansion pumping engines. The 
drawings are self-explanatory. 



^^ 



Boiler 
Pressure 




Reducing Valve 
251b. 



Trap 



Trap 



Fig. 453. Drainage System for Jackets and Receivers of Triple-expansion Pumping 

Engines. 




Feed Tank 



Fig. 454. Drainage System for Jackets and Receivers of Triple-expansion Pumping 

Engines. fl 

331. Drips under Vacuum. — Conditions frequently make it neces- 
sary to remove condensation from apparatus working under a vacuum, 
as, for example, a primary heater. 

The simplest method is to pipe the drips to the condenser and per- 
mit the condensation to gravitate to it as in Fig. 455. Where this is 
impracticable, as in an installation with the condenser above the heater, 
a steam trap is usually employed. Fig. 456 shows the apphcation of 
a Bundy trap to a vacuum or primary heater. A close-fitting weighted 
check valve W, set to open outwards, prevents intake of air through 
the discharge pipe while the trap is filhng. Connection E is made 
from the vent underneath the valve stem V back to the heater so as 
to equalize the pressures. The operation is as follows: Condensation 
gravitates from the heater through check C to the body of the trap, 



I 



SEPARATORS, TRAPS, DRAINS 



699 



the check W being closed. When the bowl is full enough to overcome 
the weight of the counterbalance, it sinks and opens up the hve-steam 
valve V. This admits steam to the trap through pipe D, which in 
turn closes check C and forces the water past the weighted check W 
to the discharge tank. After 
the water is discharged the bowl 
returns to its original position 
and closes valve V, the weight 
closes check W, the vent check 
equaUzes the pressure in the 
bowl and heater, and condensa- 
tion gravitates to the trap again. 
333. Drips under Alternate 
Pressure and Vacuum. — Occa- 
sionally the load on an engine 
is of such a character that the 
pressure in the receiver alter- 




FiG. 455. Gravity Drainage; Vacuum 
Heater. 



nates from a pressure of 30 or 40 pounds absolute to a vacuum of 
varying degree. Where the periods of vacuum operation are very few 
and of short duration, as in the average installation, no attention is 
paid to the vacuum and the condensation is removed by a trap in the 
ordinary way. If, however, the periods are of sufficient duration and 



Vacuum Heater ( | ' , 




? 




Floor Line 



Fig. 456. Method of Draining Heater under Vacuum. 

frequency, the ordinary method is not applicable and the arrange- 
ment shown in Fig. 457 may be used. The trap is placed below the 
receiver as indicated. The delivery pipe is provided with a weighted 
check or resistance valve W set so as to open outwards from the 
trap, also a spring water relief valve R. Another weighted check P 
is placed in the line leading from the vent to the atmosphere, and a 



700 



STEAM POWER PLANT ENGINEERING 



plain check C in the hne leading back into the receiver. This arrange- 
ment of valves permits the venting of the trap after discharge and 
effectually excludes air from the trap when there is less than atmos- 
pheric pressure on the receiver. With the rehef valve set to open at 




Fig. 457. Method of Draining Receiver under Alternate Vacuum and Pressure. 

a pressure in excess of the maximum receiver pressure it acts as a 
''stop" in the pipe and the water must enter the trap. When the trap 
discharges, the hve steam supphed through the pipe attached to the 
steam valve forces the water through the weighted check and rehef 
valves into the sewer or receiving tank. When working with a vacuum, 
the pressures in receiver and trap are equahzed through the vent 
connection and the condensation flows into the trap by gravity. The 
operation of discharge is the same as in the case of pressure. 

333. The Steam Loop. — Fig. 458 illustrates the principles of the 
''steam loop" for automatically returning high-pressure drips to the 
boiler. In the figure the loop is returning the condensation from a 
steam separator to a boiler above the level of the separator. The 
apparatus is very simple, consisting of a horizontal and two vertical 
lengths of plain pipe placed as indicated. Pipes R and B may be cov- 
ered but "horizontal" A is left uncovered, as its function is that of a 
condenser. The operation is as follows: Circulation is first started by 
opening stop valve at the bottom of the drop leg until steam escapes. 
The valve is then closed and the steam in the horizontal A condenses 
and gravitates to the drop leg B. On account of the slight reduction 
in pressure in the horizontal a mixture of spray and steam flows from 
the separator chamber to the horizontal, and, condensing, gravitates to 
the drop leg. The column of water in the drop leg rises until its static 
head balances the difference in pressure in the riser R and the horizontal. 
In other words, a decrease in pressure in the horizontal produces similar 
effects on the contents of the riser and drop leg but in a degree in- 



SEPARATORS, TRAPS, DRAINS 



701 



versely proportional to their densities. Any further accumulation causes 
an equal amount to pass from the bottom of the column to the boiler, 
since the pressure in the boiler is then less than that at the bottom of 
the column ; that is, the steam pressure on the top of the water column 



R. 




A\ Horizontal 



H 5f[ 



Check 



Gauge 

B 



.y.. 



0=5 



^0 



5^ 



,100^=95'^+ H 



Fig. 458. General Arrangement of the Simple ''Steam Loop." 

plus the hydrostatic head ti is greater than the pressure in the boiler. 
Once started the process is continuous and requires no further attention. 
334. The Holly Loop. — In the application of the steam loop where 
many points requiring drainage are connected to many boilers and 
conditions are more complex, some method other than the simple one 
of radiation may be advisable to secure the necessary lower pressure at 
the top of the loop. Such a method is illustrated in Fig. 459. This 
arrangement differs from the simple loop in that all condensation first 
gravitates to a ''Holly" receiver (shown in detail in Fig. 460) before 
passing into the ''riser." The receiver is placed below the lowest 
point to be drained and serves as a storage for large or unusual quan- 
tities of water and enables the riser to act at a constant rate independ- 
ent of variable discharge into the receiver. Furthermore, the lower 
pressure in the discharge chamber necessary to secure the lifting of the 
mingled steam and water through the riser, instead of being created by 
condensation as in the simple loop, is produced by a reducing valve B 
discharging into the feed-water heater. The operation of the Holly 
loop is as follows: Circulation is started by opening valve D until steam 
appears. Valve D is then closed and the reducing valve is put into 
commission. Condensation from separators, traps, and pipes gravi- 
tates to the "receiver," from which it is forced into the "riser" in the 
form of a spray. The spraying effect is produced by a series of holes 



702 



STEAM POWER PLANT ENGINEERING 




SEPARATORS, TRAPS, DRAINS 



703 



drilled in pipe A, Fig. 460. From this receiver the spray and moisture 
rise to the ''discharge chamber," on account of the lower pressure at 
that point, where the steam and entrained water are separated, the 
water gravitating to the bottom of the chamber and thence to the drop 
leg, and the steam discharging 
through the reducing valve into 
the heater. The principles of 
operation are exactly the same as 
in the simple steam loop. 

335. Returns Tank and Pump. — 
Low-pressure drips in connection 
with heating systems may be re- ^ig. 460. Holly Receiver, 

turned to the boiler along with the condensation from the heating system 
by a combined pump and receiver as shown in Fig. 461. The height of 
water in the tank controls the operation of the pump through the me- 
dium of a float and throttle valve. This combination of float and bal- 
anced throttle valve is sometimes called a "pump governor." In the 
illustration the pump forces the returns through a closed heater before 





vy////////^///y////4^/, 



» _Ii,xhaust 

2 



Engines 



Drip Trapped [I 
to Sewer — * 

Fig. 461. Returns Tank and Pump. 

delivering them to the boiler, though they are oftentimes returned di- 
rectly. The tank is vented to the atmosphere to prevent it from becom- 
ing ''air bound. " The cold-water supply or make-up water is sometimes 
discharged into the receiving tank as indicated. With open heaters the 
cold supply is ordinarily controlled by a float within the heater itself. 



704 



STEAM POWER PLANT ENGINEERING 



Air Supply 



Discharge 



336. Office Building Drains. — In the power plants of tall office build- 
ings the public sewers are often above the basement level, and it is 
necessary to remove all liquid wastes mechanically. 

The Shone pneumatic ejector has been found to serve this purpose 
effectually. This apparatus is placed in a pit in the basement floor 
into which all sewage, drips from engines, washings from boilers, and 
ground water gravitate, and are automatically discharged into the street 
sewer by means of compressed air. 

Fig. 462 gives a sectional view of a Shone ejector of ordinary con- 
struction. It consists essentially of a closed vessel furnished with inlet 

and discharge connections fitted with check 
valves, A and B, opening in opposite direc- 
tions with regard to the ejector. Two cast- 
iron bells, C and D, are linked to each other, 
in reverse positions, the rising and falling 
of which control the supply of compressed 
air through the agency of automatic valve E. 
The bells are shown in their lowest posi- 
tion, the supply of compressed air is cut ofT 
from the ejector, and the inside of the vessel 
is open to the atmosphere. The sewage 
gravitating into the ejector raises the bell C, 
which in turn actuates the automatic valve 
E, thereby closing the connection between 
the inside of the ejector and the atmosphere and opening the connection 
with the compressed air. The air pressure expels the contents through 
the bell-mouthed opening at the bottom and the discharge valve B into 
the main sewer. Discharge continues until the level falls to such a point 
that the weight of the sewage retained in the bell D is sufficient to pull it 
down, thereby reversing the automatic valve. This cuts off the supply 
of compressed air and reduces the pressure to that of the atmosphere. 
The positions of the bells are so adjusted that compressed air is not 
admitted until the ejector is full, and is not allowed to exhaust until 
emptied down to the discharge level; thus the ejector discharges a 
fixed quantity each time it operates. 

Two ejectors, each of a capacity suitable for handling the average 
flow of tributary sewage and so arranged that they can work either 
independently or together, are usually installed at each ejector station. 
The main sanitary sewer of the building usually discharges directly 
into the ejectors, the surface water, drips, etc., being collected in a 
neighboring sump. The latter is connected to the sanitary sewer 
through a trap or back-water valve. 




Fig. 462. Shone Ejector. 



CHAPTER XV 

PIPING AND PIPE FITTINGS 

337. General. — The advent of high pressures and superheat is re- 
sponsible for the ehmination of many of the older systems of piping, 
the tendency being towards greater uniformity in design, particularly 
in electric central-station work. In isolated stations the conditions of 
operation and installation are so variable that each case presents an 
entirely different problem. In any system of piping the fundamental 
object is to conduct the fluid in the safest and most economical manner. 

The material should be the best obtainable and the system so flexible 
that a break-down in one element will not necessitate the closing down 
of the entire plant. On the other hand, flexibihty increases the number 
of parts and, unless first cost is of httle importance, tends to weaken 
the system as a whole. It is a safe general proposition to say that the 
best pipe and fittings, irrespective of first cost, will prove the most 
economical in the end, but few owners of power plants are willing to 
take this view. 

338. Drawings. — An assembly drawing of the entire installation 
giving the location of all valves and fittings is necessary in order to 
avoid interference, and particularly where a number of fittings are to 
be close together. Detailed drawings should also be provided of each 
division of the piping to faciUtate installation, as, for example, the 
high-pressure steam, the exhaust steam, the feed water, the condensing 
water, the oil, the heating, and the sanitary piping. As a rule, lower 
and more uniform bids will be obtained from an isometric or perspec- 
tive sketch, as in Fig. 463, than from conventional plan and elevation 
drawings, due, no doubt, to the greater ease with which the drawing is 
interpreted. A complete set of specifications for a piping system is given 
in paragraph 479 and illustrates the usual practice along this line. 

339. Materials for Pipes and Fittings. — The following materials are 
used in the construction of pipes for steam, water, and gases. 

Average Tensile Strength. 

Low-carbon or mild steel 65,000 lb. per sq. in. 

Wrought iron 50,000 lb. per sq. in. 

Cast iron, high grade 20,000 lb. per sq. in. 

Cast steel 50,000 lb. per sq. in. 

Wrought copper 33,000 lb. per sq. in. 

Brass 18,000 lb. per sq. in. 

Special alloys and compounds 15,000-85,000 lb. per sq. in. 

705 



706 



STEAM POWER PLANT ENGINEERING 



Mild Steel. — The greater portion of the piping in the average steam 
power plant is of mild steel, lap or butt welded for high pressures and 
riveted for very low pressures and large diameters. Steel pipe is con- 
siderably cheaper than that manufactured from other material and 
fulfills practically all requirements for general service. 




FiG. 463. A Typical Isometric Pipe Drawing. 

Wrought Iron. — ^^Wrought-iron" pipe in a commercial sense refers 
to mild-steel pipe and unless stress is laid upon the term 'Spuddled iron" 
mild steel is ordinarily furnished. Puddled-iron pipe is not much in 
evidence in steam power plant work since mild steel is cheaper and 
fulfills all requirements. Wrought-iron pipe appears to resist corrosion 
to a greater extent than mild-steel pipe. Numerous laboratory in- 
vestigations have been made of late which show that mild steel is equal 
if not superior to wrought iron in many ways but in actual service the 
latter appears to have the longer life. 

Cast-iron Pipes. — Cast iron is little used for high-pressure steam 
piping except occasionally in the construction of manifold headers. 



PIPING AND PIPE FITTINGS 707 

The chief objections to cast iron for high-pressure steam are its weight 
and lack of homogeneity. It is mostly used in connection with water 
service and sanitation. For manifold headers and the Uke steel pipes 
with welded connections have superseded cast iron in the modern plant. 

Cast-steel Pipe. — Cast-steel headers are sometimes used in power 
plants for highly superheated steam, since the material is not affected 
by temperature variations to the same extent as mild steel. High first 
cost and the difficulty of securing castings free from blowholes have 
prevented its more general use. 

Copper Pipes. — Copper steam pipes were in common use for many 
years in marine service on account of their flexibility. To increase the 
bursting strength, pipes above 6 inches in diameter were generally 
wound with a close spiral of copper or composition wire. In recent 
years wrought-iron and steel pipe bends have practically superseded 
copper for flexible connections. As a rule the use of copper pipes should 
be avoided on account of the rapid deterioration of the metal under 
high temperatures and stress variations. The cost is prohibitive for 
most purposes and this alone prevents it from being seriously considered 
in the manufacture of pipe. Copper expansion joints are occasionally 
used in low-pressure work. 

Brass Pipes. — Brass is little used in the construction of pipes on 
account of its high cost. It withstands corrosive action much better 
than iron or steel and is sometimes used in connecting the feed main 
with the boiler drum. Special alloys, nickel steel, ''ferrosteel," mal- 
leable iron, and the like have been used in the manufacture of pipes, 
and possess points of superiority over wrought iron and steel for some 
purposes, but the cost is prohibitive for average steam power plant 
practice. 

Materials for Fittings. — Elbows, tees, flanges, and similar fittings 
are usually made of cast iron, malleable iron, or pressed steel, though 
cast steel, "ferrosteel," and other steel compounds are used to a limited 
extent. Standard cast-iron fittings are recommended for saturated 
steam and for pressures of 100 pounds per square inch or less, and 
extra heavy cast-iron fittings for higher pressures. Malleable-iron 
fittings are fighter and neater than cast-iron and are extensively used 
for small sizes of steam and gas pipe. Cast or pressed steel is recom- 
mended for very high pressures and superheat. 

340. Size and Strength of Commercial Pipe. — Wrought-iron and 
mild-steel pipes are marketed in standard sizes. Those most commonly 
used in steam power plants are designated as 

1. Merchant or standard pipe. 

2. Full-weight pipe. 



708 STEAM POWER PLANT ENGINEERING 

3. Large O.D. pipe. 

4. Extra heavy. 

5. Double extra heavy. 

Table 118 gives the dimensions of standard ''full- weight" pipe, 
which is specified by the nominal inside diameter up to and including 
12 inches and based on the Briggs' standard. Pipes larger than 12 
inches are designated by the actual outside diameter (O.D.), and are 
made in various weights as determined by the thickness of metal 
specified. Manufacturers specify that ''full-weight" pipe may have a 
variation of 5 per cent above or 5 per cent below the nominal or table 
weights, but merchant pipe, which is the standard pipe of commerce, 
such as manufacturers and jobbers usually carry in stock, is almost 
invariably under the nominal weight. It varies somewhat among the 
different mills, but usually lies between 5 and 10 per cent under the 
table, weight. The smaller sizes of merchant pipe, | inch to 3 inches, 
are butt-welded and the larger sizes are lap-welded. 

Extra heavy and double extra heavy pipe have the same external 
diameter as the standard, but are of greater thickness and hence the 
internal diameter is smaller. Taking the thickness of the standard 
pipe as 1, that of the extra heavy is approximately 1.4 and of the double 
extra heavy 2.8. 

Wrought-iron and steel pipes are ordinarily designed with factors of 
safety of from 6 to 15, with an average not far from 10. The standard 
hydrostatic tests to which the various pipes are subjected at the mills 
are as follows: 

Hydrostatic Pressure, 
Lb. per Sq. In. 

Standard, butt-welded, i-3 in 600 to 1,000 

Standard, lap-welded, 3-12 in 500 to 1,000 

Extra heavy, butt-welded, i-3 in 600 to 1,500 

Extra heavy, lap-welded, U'-12 in 600 to 1,500 

Double extra heavy, butt-welded, |-2| in 600 to 1,500 

Double extra heavy, lap welded, 1^-8 in 1,200 to 1,500 

The pressure necessary to burst piping is far above anything hkely 
to occur in ordinary practice on account of the thickness of material 
necessary to permit of threading. (See Table 117.) 

Riveted Pipes. — For low pressures and large diameters, pipes are 
constructed of thin sheets of boiler steel with riveted joints, the seams 
being either longitudinal and circumferential, or spiral. Such pipes 
are not necessarily limited to large sizes and low pressures, though this 
is the usual practice. 

Pipe fittings are classed as screwed, flanged or welded. 



PIPING AND PIPE FITTINGS 



709 



TABLE 117. 

BURSTING PRESSURE OF " STANDARD " MILD-STEEL PIPE.* 



No. of 


Nominal 


Actual Bursting 


No. of 
Specimen. 


Nominal 


Actual Bursting 


Diameter, 


Pressure, 


Diameter, 


Pressure, 




Inches. 


Lb. per Sq. In. 


Inches. 


Lb. per Scj. In. 


tl 


1 


7800 


1:7 


3 


3500 


t2 


1 


7700 


::8 


3 


3500 


t3 


1 


7700 
Average 7730 


t9 


3 


3000 
Average 3330 


t4 


2 


4950 


§10 


4 


1800 


t5 


2 


4800 


§11 


4 


1700 


16 


2 


5500 






Average 1750 






Average 5080 


§12 


5 


2500 








§13 


5 


2600 
Average 2550 








§14 


6 


3200 



* Tests made at Armour Institute of Technology. 

Specimens were taken at random from a lot of new pipe; length of test specimens, 5 ft. Specimens 
threaded at both ends and capped. 

t Failed at weld. t Failed in body of pipe. § Failed at threaded end. 



341. Screwed Fittings, Pipe Threads. — For screw connections the 
ends of pipes and fittings are threaded to conform to the Briggs or 
United States standard system, as shown in Fig. 464. The end of the 
pipe is tapered 1 to 32 with the axis, the angle of the thread being 




Complete Threads 
T=(0.8Di- i.8)P 



Fig. 464. Standard U. S. Pipe Thread. 

60 degrees and shghtly rounded at top and bottom. The proper length 
of perfect threads is given by the formula 

y _ (0.8 D + 4.8) ^ 
in which n 

T = length in inches, 
D = actual external diameter of the tube, inches, 



(267) 



n = number of threads per inch. 

The imperfect portion of the thread is simply incidental to the proc- 
ess of cutting. The object of the taper is to facihtate ''taking hold" 
in making up the joint. Table 118 gives the number of threads per 



710 



STEAM POWER PLANT ENGINEERING 



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spBajqx JO J9quin^N^ 


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Nominal 

Weight 

per Foot. 


4 


<MTtiiO00^«0<M«DC0t^i0O«O-«*<iOt^<N^t^OOa> 


T-l»-<(M(MCOl01>-050C<|TtioocOOOCr50iOOO 




taining 

One 

Cubic 

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i 


lO O 1-H t^ ^ (M 00 (M 00 0>(M ^ t^ 

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PIPING AND PIPE FITTINGS 



711 



inch for various sizes of standard pipe. When properly constructed a 
screwed joint will hold against any pressure consistent with the strength 
of the pipe. For example, the ultimate bursting strength of a ''stand- 
ard" 2-inch pipe is about 5000 pounds per square inch, while the 
stripping strength of the joint (with perfect threads) is 225,000 pounds. 
The threads, however, are often poorly cut and the parts screwed 
together improperly cleaned and lubricated, thus causing leakage 
between the threads. 

TABLE 119. 

STANDARD BOILER TUBES. 
Table of Standard Dimensions. 



Diameter. 


Standard 
Thickness. 


Transverse 

Areas. 


Area of Surface 

per Foot of 

Tube. 


Nominal Weight per Foot — Lb. 


1 


g 

» 


to 




c 


1 


£ 


i 


11 


Is- 


2 . 

X «c S 




2 . 
1.2 1 


W 


^ 


No. 




W 




H 


a 


i^ 


C O 




1 ^ 


Ins. 


Ins. 


Ins. 


Sq. In. 


Sq. In. 


Sq. Ft. 


Sq. Ft. 












1 


0.810 


13 


.095 


0.785 


0.515 


.262 


.212 


0.90 


1.04 


1.13 


1.24 


1.35 


u 


1.060 


13 


.095 


1.227 


0.882 


.327 


.277 


1.15 


1.33 


1.45 


1.60 


1.74 


u 


1.310 


13 


.095 


1.767 


1.348 


.392 


.343 


1.40 


1.62 


1.77 


1.96 


2.14 


n 


1.560 


13 


.095 


2.405 


1.911 


.458 


.408 


1.66 


1.91 


2.09 


2.31 


2.53 


2 


1.810 


13 


.095 


3.142 


2.573 


.523 


.474 


1.91 


2.20 


2.41 


2.67 


2.93 


2i 


2.060 


13 


.095 


3.976 


3.333 


.589 


.539 


2.16 


2.49 


2.73 


3.03 


3.32 


2i 


2.282 


12 


.109 


4.909 


4.090 


.654 


.597 


2.75 


3.05 


3.39 


3.72 


4.12 


2f 


2.532 


12 


.109 


5.940 


5.035 


.720 


.663 


3.04 


3.37 


3.74 


4.11 


4.56 


3 


2.782 


12 


.109 


7.069 


6.079 


.785 


.728 


3.33 


3.69 


4.10 


4.51 


5.00 


3^ 


3.010 


11 


.120 


8.296 


7.116 


.851 


.788 


3.96 


4.46 


4.90 


5.44 


5.90 


3^ 


3.260 


11 


.120 


9.621 


8.347 


.916 


.853 


4.28 


4.82 


5.30 


5.88 


6.38 


31 


3.510 


11 


.120 


11.045 


9.676 


.982 


.919 


4.60 


5.18 


5.69 


6.32 


6.86 


4 


3.732 


10 


.134 


12.566 


10.939 


1.047 


.977 


5.47 


6.09 


6.76 


7.34 


8.23 


^ 


4.232 


10 


.134 


15.904 


14.066 


1.178 


1.108 


6.17 


6.88 


7.64 


8.31 


9.32 


5 


4.704 


9 


.148 


19.635 


17.379 


1.309 


1.231 


7.58 


8.52 


9.27 


10.40 


11.23 


6 


5.670 


8 


.165 


28.274 


25.250 


1.571 


1.484 


10.16 


11.19 


12.57 


13.58 


14.65 



343. Flanged Fittings. — In cast-iron pipes, valves, tees, and other 
fittings the flange is always a part of the casting, but for joining the 
two ends of a steel or wrought-iron pipe the flanges may be fastened to 
the pipe in a number of ways. Fig. 465, A to H, illustrates methods 
most commonly used. In A to C the pipes are screwed into cast-iron 
or forged-steel flanges and the two faces, with metallic or composition 
gasket between, are drawn together by bolts. A illustrates the most 
common and inexpensive of flanged joints, which requires no special 
tools and can be made up at the place of erection. It gives satisfactory 
results for pressures of 100 pounds or less, but for higher pressures 
leakage is apt to take place between the threads. The flanges are 



712 



STEAM POWER PLANT ENGINEERING 



sometimes made with a long thread and a recess which can be calked 
with soft metal. A similar joint is made with the pipe screwed be- 
yond the face of the flange and the two faced off together, either plane 
or as shown in B, which is known as a male and female or hydraulic 



Smooth Face 




Screwed 



Raised Face 

TZT 



Welded 



Shrunk 




Screwed 




Eiveted 



rf 'Tg- fcl ilH fr l 




.Screwed & Peened 




Fig. 465. Types of Pipe Flanges. 



joint. This method forms a very rehable joint, since the ends of the 
pipe bear on the gasket, and the gasket is prevented from being blown 
out. An objection lies in the difficulty of opening the line to remove 
the gasket or replace a fitting. C is a modification known as the 
tongued and grooved joint, which uses an extremely narrow gasket. 
Such flanges may be subjected to severe strains when the bolts are 
drawn up, owing to the small area of contact. Corrugated copper or 
steel gaskets are recommended, since soft material is apt to be squeezed 
out. In C the ends of the pipe are peened, which is an improvement 
over the simple screwed joint. D illustrates a shrunk joint. The 
flanges are bored for a shrink fit and forced over the pipe when at a 
red heat. After cooling the end is beaded over into a recess on the face 
of the flange and a light cut taken from both. H shows a modifica- 
tion in which the hub is riveted to the pipe. E illustrates a joint con- 
structed by rolling the pipe into a corrugation in the flange. The end 
of the pipe is then faced off flush. 



PIPING AND PIPE FITTINGS 



713 



TABLE 120. 

DIMENSIONS OF CAST-IRON PIPE. 





Standard Thickness and Weight. 




Class A. 




Class B. 




Class C. 


Nominal 


100 Feet Head. 




200 Feet Head. 


300 Feet Head. 


Inside 


43 Pounds Pressure. 


86 Pounds Pressure. 


130 Pounds Pressure. 


Diam- 
















eter, 
Inches. 


Thick- 


Weight per 


Thick- 


Weight per 


Thick- 


Weight per 




ness, 
Inches. 


Foot. 


Length. 


ness, 
Inches. 


Foot. 


Length. 


ness, 
Inches. 


Foot. 


Length. 


4 


.42 


20.0 


240 


.45 


21.7 


260 


.48 


23.3 


280 


6 


.44 


30.8 


370 


.48 


33.3 


400 


.51 


35.8 


430 


8 


.46 


42.9 


515 


.51 


47.5 


570 


.56 


52.1 


625 


10 


.50 


57.1 


685 


.57 


63.8 


765 


.62 


70.8 


850 


12 


.54 


72.5 


870 


.62 


82.1 


985 


.68 


91.7 


1,100 


14 


.57 


89.6 


1,075 


.66 


102.5 


1,230 


.74 


116.7 


1,400 


16 


.60 


108.3 


1,300 


.70 


125.0 


1,500 


.80 


143.8 


1,725 


18 


.64 


129.2 


1,550 


.75 


150.0 


1,800 


.87 


175.0 


2,100 


20 


.67 


150.0 


1,800 


.80 


175.0 


2,100 


.92 


208.3 


2,500 


24 


.76 


204.2 


2,450 


.89 


233.3 


2,800 


1.04 


279.2 


3,350 


30 


.88 


291.7 


3,500 


1.03 


333.3 


4,000 


1.20 


400.0 


4,800 


36 


.99 


391.7 


4,700 


1.15 


454.2 


5,450 


1.36 


545.8 


6,550 


42 


1.10 


512.5 


6,150 


1.28 


591.7 


7,100 


1.54 


716.7 


8,600 


48 


1.26 


666.7 


8,000 


1.42 


750.0 


9,000 


1.71 


908.3 


10,900 


54 


1.35 


800.0 


9,600 


1.55 


933.3 


11,200 


1.90 


1141.7 


13,700 


60 


1.39 


916.7 


11,000 


1.67 


1104.2 


13,250 


2.00 


1341.7 


16,100 


72 


1.62 


1283.4 


15,400 


1.95 


1545.8 


18,550 


2.39 


1904.2 


22,850 


84 


1.72 


1633.4 


19,600 


2.22 


2104.2 


25,250 










Ot: 









* Adopted standards of Am. Water W'ks Ass'n. The above weights are per length to lay 12 feet, includ- 
ing standard sockets; proportionate allowance to be made for any variation. All weights are approximate. 

Dimensions of Riveted Steel Pipes: Power, March 7, 1911, p. 377. 

One of the best commercial joints is illustrated by F and is known 
as the lap joint. The pipe is expanded as indicated and a light cut is 
then taken from the flared ends to insure a tight joint. The flanges 
are loose and permit of considerable flexibility in shifting them through 
various angles. This is sometimes called the Van Stone joint. 

Pipes with flanges welded on the end as in G have proved the most 
reliable of all and though costly are considered the standard for high- 
pressure and high-temperature work. The faces are ordinarily raised 
^V to yV inch inside the bolt holes and ground to a steam-tight fit, so 
that thick gaskets are unnecessary. 

For moderately high pressures and temperatures any of the joints 
when well made will prove satisfactory. For extreme^ high pres- 
sures and temperatures the lap or welded joints are preferable. 



714 



STEAM POWER PLANT ENGINEERING 



13. 



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PIPING AND PIPE FITTINGS 



715 



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rt _ ^ ^ ^ _i CMCMCMCM' 



716 



STEAM POWER PLANT ENGINEERING 



In 



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PIPING AND PIPE FITTINGS 



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23 



718 



STEAM POWER PLANT ENGINEERING 



Corrugated steel gaskets covering the entire annular area inside the 
bolt holes are highly satisfactory for high pressures and tempera- 
tures. In a number of recent plants the tips of the flanges are welded 
by an oxy-acetylene torch to insure tightness. 
See Fig. 466. 

The comparative costs of various flanges are 
given in Table 123. 

Tables 121 and 122 give the dimensions of 
standard and extra heavy fittings as adopted by 
a joint committee of the manufacturers and of 
the American Society of Mechanical Engineers. 
This new schedule, ''The American Standard of 
1914," went into effect January 1, 1914. 

The following explanatory notes refer to Tables 
121 and 122: 




Fig. 466. Pipe Flange 
with Welded Tip. 



(a) Standard and extra heavy reducing elbows carry same dimen- 
sions center to face as regular elbows of largest straight size. 

(h) Standard and extra heavy tees, crosses and laterals, reducing on 
run only, carry same dimensions face to face as largest straight size. 

(c) If flanged fittings for lower working pressure than 125 pounds 
are made, they shall conform in all dimensions, except thickness of 
shell, to this standard and shall have the guaranteed working pressure 
cast on each fitting. Flanges for these fittings must be standard 
dimensions. 

(d) Where long radius fittings are specified, it has reference only to 
elbows which are made in two center-to-face dimensions and to be 
known as elbows and long radius elbows, the latter being used only 
when so specified. 

(e) All standard weight fittings must be guaranteed for 125 pounds 
working pressure, and extra heavy fittings for 250-pound working pres- 
sure, and each fitting must have some mark cast on it indicating the 
maker and guaranteed working steam pressure. 

(/) All extra heavy fittings and flanges to have a raised surface of 
xV iiich high inside of bolt holes for gaskets. 

Standard weight fittings and flanges to be plain faced. 

Bolt holes to be | inch larger in diameter than bolts. 

Bolt holes to straddle center line. 

(g) Size of all fittings scheduled indicates inside diameter of ports, 
except for heavy fittings 14 inches and larger when the port diam- 
eter is f inch smaller than nominal size. 

(h) The face-to-face dimension of reducers, either straight or eccen- 
tric, for all pressures, shall be the same face to face as given in table of 
dimensions. 

(i) Square head bolts with hexagonal nuts are recommended. 

For bolts If inch diameter and larger, studs with a nut on each end 
are satisfactory. 

Hexagonal nuts for pipe sizes 1 inch to 46 inches on 125-pound stand- 



PIPING AND PIPE FITTINGS 



719 



ard and 1 inch to 16 inches on 250-pound standard can be conveniently 
pulled up with open wrenches of minimum design of heads. Hexag- 
onal nuts for pipe sizes 48 inches to 100 inches on 125-pound and 18 
inches to 48 inches on 250-pound standards can be conveniently pulled 
up with box or socket wrenches. 

(j) Twin elbows, whether straight or reducing, carry same dimen- 
sions center to face and face to face as regular straight size ells and tees. 
I Side outlet elbows and side outlet tees, whether straight or reducing 
sizes, carry same dimensions center to face and face to face as regular 
tees having same reductions. 

' (k) Bull head tees or tees increasing on outlet will have same center- 
to-face and face-to-face dimensions as a straight fitting of the size of the 
outlet. 

(Z) Tees and crosses 9 inches and down, reducing on the outlet, use 
the same dimensions as straight sizes of the larger port. 

Sizes 10 inches and up, reducing on the outlet, are made in two 
lengths depending on the size of the outlet as given in the table of 
dimensions. 

Laterals 3J inches and down, reducing on the branch, use the same 
dimensions as straight sizes of the larger port. 

(m) Sizes 4 inches and up, reducing on the branch, are made in two 
lengths depending on the size of the branch as given in the table of 
dimensions. 

The dimensions of reducing flanged fittings are always regulated by 
the reductions of the outlet or branch. Fittings reducing on the run 
only, the long body pattern will always be used. 

Y's are special and are made to suit conditions. 

Double sweep tees are not made reducing on the run, 

(n) Steel flanges, flttings and valves are recommended for superheated 
steam. 



TABLE 123. 

COMPARATIVE COST OF VARIOUS PIPE FLANGE FITTINGS, 
(Circular from the Crane Company.) 



12-INCH PIPE. 



, 


i 


a 


33 


il 




13 


i 
1 


J 

IS 


Cast iron 


i 7.40 

8.70 

9.90 

22.40 

26.40 


S16.00 
18.40 


$18.00 
20.00 








$13.00 
16.00 
18.00 
25.00 
30.00 


$21.00 


Ferrosteel 








23.40 


Malleable iron 


$22.00 








Cast steel 


28.40 
32.40 


34.00 
38.00 


$33!6o 
37.00 


$4i!6o 


33 40 


Weldless steel 


37.40 







Any of the above screwed, shrunk, welded, rolled, or single-riveted flanges can be 
furnished with male or female face at $1.25 extra. 

The screwed or welded flanges can be furnished with tongued or grooved face at 
$1.25 extra. 

Any of the above screwed, shrunk, or single-riveted flanges can be furnished with 
calking recess at $1.25 extra. 



720 



STEAM POWER PLANT ENGINEERING 



In modern high-temperature, high-pressure practice all nozzles for 
connecting the leads are welded to the headers thereby insuring a 
minimum number of joints. 

343. Loss of Heat from Bare and Covered Pipe. — Steam pipes, feed- 
water pipes, boiler steam drums, receivers, separators and the like 
should be covered with heat-insulating rriaterial to reduce heat losses 
to a minimum. By properly applying any good commercial covering 



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50 100 150 200 250 300 350 400 450 

Temperature Difference— Degrrees Fahrenlieit 

(Pipe Temp.— Room Temp.) 

Fig. 467. Total Losses from Bare Pipe 



500 



from 75 per cent to 95 per cent of the heat loss may be prevented. 
Numerous investigations have been made relative to the heat losses 
from bare and covered pipes, but the results have been far from har- 
monious. The most trustworthy results appear to be those based upon 
the investigations of L. B. McMillan (Trans. A.S.M.E., Vol. 38, 1916). 
The loss of heat from bare pipes, as found by McMillan, is given in 
the curves of Fig. 467 and the insulating properties of a number of 
well-known pipe coverings are shown in Fig. 468. From the curved 



PIPING AMJ) PIPE FITTLXCIS 



721 



in Fig. 467 it will ])e seen that heat loss from bare pipes is so great 
that covering will pay for itself in a comparatively short time. The 
curves, Fig. 469, showing the relation of the rate of loss per sq. ft. of 
covering surface to the temperature difference between the covering 



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50 100 150 300 250 300 a^O 400 
Temperature Difference, De^Tecs Fahrenlieit 
(Pipe Temp.-Room Temp.) 



450 500 



Fig. 468. Heat Loss Through Pipe Covering (Singk^ Thickness).* 

surface and the surrounding air is one of the most important results 
obtained by McMillan and furnishes the data required for calculating 
the heat loss from covered pipe having its surface finished in white 
canvas; thus, for fincUng the heat loss through any thickness of any 



* The average thicknesses in inches of the coverings, Fig. 468, are as follows : 1 — 
1.08; 2 — 1.12; 3 — 0.96; 4—1.04; 5 — 1.25; 6—1.10; 7—1.07; 8 — 0.99; 9 — 
1.00; 10 — 0.96; 11 — 1.10; 12 — I.IG; 13 — 1.16; 14 — 0.99; 15—1.16; 16 — 1.10; 
17—1.00; 19—1.05; 24 — 0.95. 



722 



STEAM POWER PLANT ENGINEERING 



material of which the conductivity is known, at any temperature 
difference between the pipe and room up to 500 deg. fahr. : 

k{ti- t - d) 



H2 = 



r2(l0ger2 — logen) 



n 



H2 = — Hi, 

r2 



(268) 

(269) 
m which ^2 

H2 = heat loss per sq. ft. of outside covering surface, B.t.u. per hr., 
k = conductivity of the material, B.t.u. per hr. per sq. ft. per in. 
thickness per degree temperature difference, 
^1 and t = temperatures, respectively, of the pipe and of the air in 

the room, deg. fahr., 
r2 and n = radii, respectively, of the outer and inner surfaces of the 
covering, in., 
d = temperature difference between the covering and air corre- 
sponding to a rate of loss H2, 
Hi = heat loss per sq. ft. of pipe surface, B.t.u. per hr. 



16.0 

15.0 

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n 



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13.0 
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to 

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86.0 

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1 4.0 
"§3.0 

2.0 

1.0 

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y. 


































y 


/ 


































/ 


^ 


































/ 


< 


































/ 


/ 


































y 


/- 


































A 


r 




































/ 




































/ 




































/ 


/ 




































/ 







































40 80 130 160 200 240 280 320 380 

s per Square Foot of Outer Surface of Covering. B.T.U. (=H 2) 



Fig. 469. Relation of Heat Losses to Temperature. 
The conductivity may be calculated as follows: 



k = 



HiVi (loge r2 - loge ri) 



in which ^i ~ ^2 

^ = temperature of the outer surface of the covering, deg. fahr. 

Other notations as in equations (268) and (269). 
These laws are best illustrated by examples 66 and 67. 



(270) 



PIPING AND PIPE FITTINGS 723 

Example 66. A steam pipe 5.6 in. outside cliaiiicter is covered with 
single-thickness J-M 85 per cent magnesia, 1.13 in. thick, temperature 
of the pipe 380 deg. fahr., room temperature 80 deg. fahr. Required 
the conductivity per inch thickness for the given conditions. 

From Fig. 468 the rate of heat loss per hour per sq. ft. per deg. tem- 
perature difference is 0.455 B.t.u. Therefore, Hi = 300 X 0.455 = 136.5 

and Hi = 136.5 X ^ ^ (^ + I.I3] = 97.2 B.t.u. From Fig. 469 the 

temperature difference between outer covering surface and air corre- 
sponding to a loss of 97.2 B.t.u. is 65 deg. fahr. Therefore, the 
temperature difference between inner and outer covering surfaces is 
300 — 65 = 235 deg. fahr. Substituting these values in equation (270) 
and solving for k, 

, _ 136.5 X 2.8 (log. 3.93 - log^ 2.8) _ „ __^ 
k - 235 ~ ^•^^^• 

Example 67. If the pipe in Example 66 is covered with 3-inch thick- 
ness of material, other conditions remaining the same, calculate the 
heat loss per sq. ft. of pipe surface per hr. per degree temperature dif^ 
ference. 

From equation (268) 

^ 0.551 (380 - 80 - ^) 
^' (2.8 + 3) (log. 5.8 -log. 2.8) 
= 0.13 (300 - d). 

Now assume d = 20 deg. Then H2 from Fig. 469 = 25.5 B.t.u. 
But H2 from equation (268) = 0.13 (300 - 20) = 36.4. This shows 
that d must be greater than 20. Assume d = 30. Then H2 from Fig. 
469 = 39.5 B.t.u. and from equation (269) H2 = 0.13 (300 - 30) 
= 35.1. This shows that d must be less than 30. By cut and trial 
the correct value ^2 = 27 may be obtained. Then H2 = 0.13 X (300 
— 27) = 35.5. Substitute this value of H2 in equation (269) and solve 

^'''^" o. X 2.8 ^ 

35.5 = --X Hi, 

0.0 

from which Hi = 73.5 B.t.u. per hr. per sq. ft. Loss per sq. ft. per 
hr. per deg. temperature difference between the pipe surface and air 
in the room = 73.5 ^ 300 = 0.245 B.t.u. 

Pipe covering is applied in sections molded to the required forn and 
held to the pipe by bands, or may be applied in a plastic form. The 
former is more readily applied and removed, and is usually adopted for 
pipes, while the valves and fittings are generally covered with plastic 
material. Piping should be tested under pressure before being covered, 
since leaks destroy the efficiency and life of the covering. If the sur- 
rounding atmosphere is moist the covering should be given two or three 
coats of good paint. Coverings are sometimes applied to cold water 
pipe to prevent sweating. 

Identification of Power House Piping bij Colors : Power and Engineer, April 26, 
1910, p. 752. 



724 



STEAM POWER PLANT ENGINEERING 



TABLE 124. 

COEFFICIENTS OF LINEAR EXPANSION PIPING MATERIALS. 



Material. 



Wrought iron and mild steel... 

Wrought iron 

Cast iron 

Cast steel 

Hardened steel 

Nickel-steel, 36 per cent Nickel 

Copper, cast 

Copper, wrought 

Lead 

Cast brass 

Brass wire and sheets 

Tin cast 

Tin hammered 

Zinc cast 

Zinc hammered 



Temperature 
Range. 



32-212 
32-572 
32-212 
32-212 
32-212 
32-572 
32-212 
32-572 
32-212 
32-212 
32-212 
32-212 
32-212 
32-212 
32-212 



Mean Coeffi- 
cient per De- 
gree F. 



0.00000656 
0.00000895 
0.00000618 
0.00000600 
0.00000689 
0.00000030 
0.00000955 
0.00001092 
0.00001580 
0.00001043 
0.00001075 
0.00001207 
0.00001500 
0.00001633 
0.00001722 



LINEAR EXPANSION OR CONTRACTION OF CAST IRON IN INCHES PER 
100 FEET, — DEGREES F. 



Temperature Difference. 


Expansion. 


Temperature Difference. 


Expansion. 


100 
150 
200 
250 


0.72 
1.1016 
1.5024 
1.9260 


300 
400 
500 
600 
800 


2.376 
3.360 
4.440 
5.616 
7.872 









Multiply by 1 .1 for wrought mild steeL 
Multiply by 1 .5 for wrought copper. 
Multiply by 1.6 for wrought brass. 

344. Expansion. — One of the most difficult problems in the design 
of a piping system is the proper provision for expansion and contraction 
due to change in temperature. If a pipe is immovably fixed at both 
ends and under no strain when cold, and the temperature is increased, 
as by the admission of steam, it is subjected to a compression propor- 
tional to the rise in temperature (within the elastic limit). The axial 
force exerted due to expansion may be expressed 

P = EA {h - t) tx (Mechanics of Engng., Church, p. 218), (271) 

P = force in pounds, 

E = modulus of elasticity (average for steel pipe = 30,000,000), 

h = final temperature, deg. fahr. (the temperature of the pipe is 
practically that of the steam), 



PIPING AND PIPE FITTINGS 



725 



t = initial temperature, 
/z = coefficient of expansion, 
A = sectional area of the pipe material, sq. in. 

Example 68. A 6-inch standard extra heavy steel iron pipe 200 feet 
long at 66 deg. fahr., heated to 366 deg. fahr. (the temperature cor- 
responding to steam at 165 pounds per square inch absolute pressure), 
required the axial force exerted. 

Here 

E = 30,000,000; h = 366; ^ = 66; = 0.000007 (approx.), 
A = 8.5 sq. in. 

Substituting these values in equation (271), 

P = 30,000,000 X 8.5 (366 - 66) 0.000007 
= 535,500 lb. 

Unless well braced throughout its entire length the pipe will buckle 
and become distorted. If free to expand its length would increase. 
The total increase in length is the sum of the elongation due to pressure 
and that due to increase in temperature. The increase in length due 
to pressure is negligible except for extremely high pressures and long 
lengths of thin pipe, but that due to temperature may be considerable. 

TABLE 125. 



SAFE 


EXPANSION VALUES 01 


^ 90-DEGREE WROUGHT 


STEEL BENDS IN INCHES. 


(Full weight or extra heavy pipe.) 










Mean Radius of Bend (in Inches). 




























12 


15 


20 


30 


40 


50 


60 


70 


80 


90 


100 


110 


120 


1 


1 

4 


t 


f 


u 


3^ 




















2 


i 


f 


i 


1 


n 


2^ 


3^ 


5^ 






... 










2h 




i 




7 


u 


2i 


3i 


^ 


5f 














3 




i 


1 


f 


n 


U 


3^ 


3^ 


4| 


6 












3 








1 


1 


U 


n 


3^ 


H 


5? 












4 








^ 


1 


H 


2 


2| 


'Si 


4? 


5f 










U 










1 


n 


n 


2^ 


31 


44 


5i 










5 








1 


3 
4 


n 


n 


2i 


3 


3? 


4| 


5f 






6 








f 


1 


1 


n 


11 


22 


3« 


3| 


4| 


5t 


8 










4 


i 


1 


u 


1h 


'M 


3 


3| 


4t 


10 














I 


n 


H 


2 


2^ 


21 


3^ 


12 














3 
4 


1 


n 


11 


2 


2h 


2| 


14 
















7 
8 


n 


n 


1- 


2\ 


2k 


15 


















I 


n 


1| 


2 


2% 


16 


















I 


u 


li 


n 


2i 


18 
























n 


U 


20 


















. . . 








u 



For any compound expansion bend multiply the tabular value by the number of 
90-degree bends, thus for a " U " bend multiply the tabular values by 2; for an " ex- 
pansion U " bend multiply by 4. 



726 



STEAM POWER PLANT ENGINEERING 



The increase in length for both conditions may be expressed 

_ paL 

lt = n ih - t) L, 



(272) 
(273) 



in which 

Ip = increase in length due to the internal pressure, in., 
It = increase in length due to the temperature difference, 
p = boiler pressure, lb. per sq. in. gauge, 
a = inside area of the pipe, sq. in., 
L = length of the pipe, in. 

Other notations as in equation (271). 

Example 69. A 12-inch extra-heavy steel pipe is 100 feet long when 
cold (70 deg. fahr.): Required the increase in length when carrying 
superheated steam at 250-lb. gauge pressure, temperature 670 deg. fahr. 

Here p = 250, a = 108.4, L = 1200, E = 30,000,000, A = 19.25, 
t^ = 670, i = 70, M = 0.0000075 (the coefficient of Unear expansion is 
known to increase with the temperature; the value assumed here is a 
purely arbitrary one). 

Substituting these values in equations (272) and (273), 
250 X 108.4 X 1200 



U = 



0.056 in., which is negligible, 



30,000,000 X 19.25 
It = 0.0000075 (670-70) 1200 = 5.4. 

Since the forces produced by expansion are practically irresistible 
the pipe is invariably allowed to expand and its movement is prevented 
from unduly stressing the fittings and connections by 

1. Long radius bends. 

2. Double-swing screwed fittings. 

3. Expansion joints. 

TABLE 126. 

MINIMUM DIMENSIONS FOR PIPE BENDS. 





Radius of Bend, In. 






Radius of Bend, In. 










Lengths of 






Lengths of 


Size of 






Straight Pipe 


Size of 






Straight Pipe 


. Pipe, In. 


Full Weight 
Pipe. 


Extra 
Heavy 
Pipe. 


on Each 
Bend, In. 


Pipe, In. 


Full Weight 
Pipe.^ 


Extra 
Heavy 
Pipe. 


on Each 
Bend, In. 


2h 


12.5 


7 


4 


8 


40 


28 


9 


3 


15.0 


8 


4 


9 


45 


35 


11 


31 


17.5 


10 


5 


10 


50 


40 


12 


4 


20.0 


12 


5 


12 


60 


50 


14 


4| 


22.5 


14 


6 


14 


70 


65 


16 


5 


25.0 


15 


6 


15 


75 


70 


16 


6 


30.0 


20 


7 


16 


80 


78 


18 


7 


35.0 


24 


8 


18 


108 


88 


18 



PIPING AND PIPE FITTINGS 



727 



Where practical long radius bends will prove most satisfactory.* 
Fig. 470 shows a number of standard bends and Table 126 gives the 
minimum radii and lengths of straight pipe at the end of each bend as 
recommended by the Crane Company. The amount of expansion 




single: OFFSET 
U BEND 



F.XPAN3ION U BEND 




EXPANSION LOOP 



EXPANSION U BEND 



CROSS OVER 



Fig. 470. Types of Expansion Bends. 

absorbed by a standard 90-degree quarter bend and other shapes may 
be taken from Table 125. 

Figs. 484 and 485 show appHcations of pipe bends to boiler and 
header connections. 

*At the Essex Power Station of the Public Service Electric Co. of New Jersey 
there are no expansion joints in the headers. The headers are installed under a 
tension between anchorages, which causes an elongation equal to about one-half 
the expansion of the section from normal temperature to that of the steam. When 
the headers are at ordinary room temperature they are in tension, and when at the 
temperature of the steam they are in compression. 



728 



STEAM POWER PLANT ENGINEERING 




Fig. 471 shows a double-swing screwed joint in which expansion 
causes the fittings to turn slightly and thus reheve the strain. This 
method is usually adopted where long radius 
bends are not practicable on account of lack 
of space and where screwed fittings are used. ^ 

Slip joints, Fig. 472, are now little used ex- 
cept with very large pipes and where space 
prohibits long radius bends. When sUp joints 
are employed the pipe must be securely an- 
chored to prevent the steam pressure from 
forcing the joint apart and at the same time 
permit the pipe in expanding to work freely 
in the stuffing box. Sagging of the pipe on 
either side, which might cause binding in the 
joint, is prevented by suitable supports. 

Elasticity and Endurance of Steam Piping: Power, 
Feb. 23, 1915, p. 278. 



Fig. 470a. U Bends for 
Large Headers when 
Overhead Space is 
Limited. 




UB 



FRONT ELEVATION 



Fig. 471. ''Double-swing" 
Expansion Joint. 




Fig. 472. Slip Expansion Joint. 



345. Pipe Supports and Anchors. — Pipe fines must 
be supported to guard against excessive deflection and 
vibration. Supports are conveniently classified as (1) 
hangers, (2) wall brackets, and (3) floor stands. 

Fig. 473 illustrates a type of hanger for suspending 
pipes from I beams. The supports being free to swing, 
no provision for expansion is necessary. A properly 
designed hanger may be readily removed without dis- 
turbing the pipe line, and should be adjustable to 
facifitate ''lining up." If of rigid construction the 
lower end should be provided with a roller. 

Fig. 474 gives the details of a wall bracket with rolls and roll binder. 
Supports adjacent to long radius bends should be provided with roll 
binders as illustrated to prevent the pipe from springing laterally, 




Fig. 473. 

A Typical Pipe 

Hanger. 



PIPING AND PIPE FITTINGS 



729 





i^ 




Fig. 474. A Typical Wall Fig. 475. A Typ- 
Bracket with Binding Roll. ical Floor Stand, 



Fig. 476. A Typical Pipe 
Anchor. 




Fig. 477. Method of Suspending and Counterbalancing Expansion Loops in 

Steam Mains. 



730 



STEAM POWER PLANT ENGINEERING 



but they may otherwise be omitted. The rollers are often made 
adjustable to facilitate Uning up. 

Fig. 475 illustrates a typical floor stand. Pipe lines are usually 
securely anchored at suitable points in a manner similar to that illus- 
trated in Fig. 476, the pipe resting on a saddle and being rigidly clamped 
to the bracket by a flat iron band with ends threaded and bolted. 
This limits expansion to one direction and prevents excessive strain on 
the fittings. 

Fig. 477 illustrates a method of suspending and counterbalancing 
expansion loops in a main header and Fig. 478 a flexible support for a 
large vertical exhaust header. 




Fig. 478. Spring Support for 30-inch Exhaust Pipe. 



346. General Arrangement of High-pressure Steam Piping. — The gen- 
eral arrangement of piping depends in a great measure upon the space 
available for engines and boilers. 

The engine and boiler room may be placed 

(1) Back to back, 480. 

(2) End to end, 479. 

(3) Double decked, 486. 

The hack-to-back arrangement is the most common and, other things 
permitting, is to be preferred on account of the short and direct con- 
nection between prime movers and boilers and the ease of enlargement. 
The engine and boiler rooms are separated by a wall, and as much of 
the piping as possible is located in the boiler room. 



PIPING AND PIPE FITTINGS 



731 



The end-to-end arrangement is ordinarily limited to situations where 
the distribution of space precludes the back-to-back system. 

The double-decked arrangement 
is frequently used where ground 
space is Hmited or expensive. 

Prime movers and boilers are 
connected in a variety of ways 
through steam headers as shown 
in Figs. 479 to 489: 

1. Spider system, Fig. 480. 

2. Single header, Fig. 481. 

3. Duplicate system. 

4. Loop or ring header. Fig, 483. 

5. The "unit" system. Fig. 484. 
The spider system is often used 

in small plants. In this arrange- 
ment all branch pipes are brought 
to one central header which is 
made as short as possible. The 
shortness of such a header mini- 
mizes danger from breakdowns, 
and brings all the principal valves 
close together. 

The single-header system is per- 
haps the most common, since it 
embodies simplicity, low first cost, 
and provision for extension. 

The duplicate system is losing 
favor, since experience shows that 
the extra cost of the duplicate 
mains will usually give better re- 
turns in continuity of operation 
and maintenance if invested in 
high-grade fittings on a single-pipe 
system. A small auxihary header 
is occasionally used in plants 
where double mains are desired. 
In the new River Station of the 
Buffalo General Electric Company 
the steam main is in duplicate, 
see Figs. 487 and 488, but this arrangement is for the purpose of 
insuring flexibility and for keeping down the size of pipe and not 




bD 

'a 

S 



732 



STEAM POWER PLANT ENGINEERING 



as a protection against breakdown. Both headers are in use simul- 
taneously. 

The loop header is well adapted where a large number of steam 
engines, elevator pumps, air compressors, and miscellaneous steam- 
consuming appliances are crowded together in a comparatively small 
space. 

Large modern power plants are, by the latest practice, divided into 
complete and independent units, as in Fig. 484, each prime mover 



Battery No.l 




-^^^^^^^^^^^^^^.^^^^^^-^^^^^:?^ 



Battery No.2 



Battery No.3 



3.2 No.3 No.1 

Fig. 480. "Spider" System. 



having its own boiler equipment, coal and ash-handling machinery, 
feed pumps, and piping, operated independently of the rest of the 
plant. 

The steam mains are usually cross connected so that steam from any 
boiler unit may be led to the adjacent prime mover. 

Figs. 484 and 485 show the general arrangement of the steam piping 
at the Yonkers Power House of the New York Central, illustrating a 
typical ''unit system." The turbines are connected in pairs by 14- 
inch loops, each turbine taking steam from either of two banks of 
four boilers. The high-pressure steam piping is of mild steel with 
modified reinforced ''Van Stone" joints. The high-pressure valves 
are of the split-disk pattern with semi-steel bodies. Expansion is 
taken up by the long sweep bends. 

Plants using superheated steam are sometimes piped to supply 
saturated steam to the auxiliaries as illustrated in Fig. 489. The 
boiler branch E, leading to the main header, normally supplies super- 



PIPING AND PIPE FITTINGS 



733 




734 



STEAM POWER PLANT ENGINEERING 



heated steam to the engines. C is an auxihary main supplying the 
air pumps, stoker engines, and other auxiliaries with saturated steam 

from branch pipe D. 

347. Size of Steam Mains. — 
Until quite recently it was the 
usual practice to employ a com- 
mon header running the entire 
length of the plant and to con- 
nect all boiler and engine leads 
with this header. With the low 
steam velocities used at that 
time headers as large as 24 
inches in diameter were not un- 
common. This type of station 
is rarely built at the present 
day except, perhaps, for very 
small plants. In the various 
large power houses recently 
built in this country with ulti- 
mate capacities of from 100,000 
to 250,000 kilowatts, the largest 
steam headers are not over 18 
inches in diameter. In some 
recent designs the pipes leading 
from the header to the engines 
are two sizes smaller than called 
for by the engine builders. In 
this case large receiver separa- 
tors two to four times the 
volume of the high-pressure 
cylinder are provided near the 
throttle. The pipes between 
receiver and engine are full size. 
The object of the arrangement 
is to give (1) a constant flow 
of steam, (2) a full supply of 
steam close to the throttle, 
Fig. 482. Typical Auxiliary Header System, ^nd (3) a cushion near the en- 
gine for absorbing the shock caused by cut-off. 

With saturated steam and boiler pressures from 125 to 150 pounds 
a maximum velocity of 8000 feet per minute is allowed in the main 
and as high as 9000 feet per minute between header and receiver. 




PIPING AND PIPE FITTINGS 



735 




736 



STEAM POWER PLANT ENGINEERING 




I 



PIPING AND PIPE FITTINGS 



737 



|^3§>^A- 



9 ! 8 Automatlo 

' NoD-R«turD ValTO 




Detail Plan 



(Power) 



Fig. 485. Details of Boiler Steam Piping, Yonkers Power House of the 
New York Central R.R. 



738 



STEAM POWER PLANT ENGINEERING 



With steam turbines using highly superheated steam velocities as high 
as 16,000 feet per minute have been allowed during peak loads but 
the pressure drop between boiler and prime mover is apt to be ex- 




FiG. 486. Typical Double-deck Boiler Installation (New York Steam Co.). 



cessive. Exhaust steam velocities range from 12,000 to 36,000 ft. per 
minute, depending upon the pressure of the steam and the length of 
the piping. 



PIPING AND PIPE FITTINGS 



739 




i 

a 

o 



W 



m 



03 

m 

•I 

'a, 

a 

03 

(U 



W 



740 



STEAM POWER PLANT ENGINEERING 



348. Flow of Steam in Pipes. — In designing a piping system the 
engineer is chiefly concerned with the size of pipe which will deliver 
a given weight of steam under given initial conditions to a distant 
point at a predetermined pressure drop. In small plants extreme ac- 
curacy in determining the size of pipe is not necessary; it is better 
to err in the installation of too large a pipe than one too small. In 
large stations where the pipes are large and the pressure is high the 
cost of piping increases rapidly with the size and greater accuracy is 
essential. Since the weight of steam discharged through any system 




Unit No. 1 



Unit Na 



Fig. 488. Plan of Main Steam Piping, River Station, Buffalo General 
Electric Company. 



of piping is a direct function of the drop in pressure it is evident that 
the greater the drop the larger will be the weight discharged. A large 
drop in pressure permits of a smaller pipe and lower radiation losses, 
but a point is soon reached where the economy in the size of pipe is 
more than offset by the loss in available energy due to the reduced 
pressure at the point of appUcation. There seems to be no fixed rule 
for determining the drop most suitable for any given set of conditions. 
In reciprocating engine practice involving the use of saturated steam 
and in which the pipe leads directly to the inlet nozzle the maximum 
drop in pressure ordinarily varies from J to 1| pounds per hundred 
feet of pipe, corresponding to a maximum velocity of approximately 



PIPING AND PIPP] FITTINGS 



741 



6000 feet per minute. In a iiuinber of installations in which a hirge 
receiver is placed next to the inlet nozzle pressure drops of 1.5 to 2.5 
pounds per 100 feet of pipe with corresponding maximum velocity of 
about 9000 f(^et per minute have given satisfactory results. For very 
long pipe lines the pressure drop per 100 feet must necessarily be small 
in order to avoid low pressures at the point of delivery. In steam 
turbine practice involving the use of high pressure and superheat 
pressure drops as high as 3.5 pounds per 100 feet of pipe have been 
allowed during periods of maximum discharge. Under the latter con- 
ditions pipe velocities as high as 16,000 feet per minute have been 




Fig. 489. Overhead Boiler Piping, Quiiicy Point Power Plant of the Old Colony 
St. Ry. Co., Quincy Point, Mass. 



obtained. It must be remembered that the pressure drop through the 
piping is but a small portion of the total drop from boiler to prime 
mover because of the additional resistance of the dry pipe, super- 
heater, valves, and fittings; consequently large pressure drops through 
the piping alone may cause excessive drops from boiler to prime mover. 
(See following paragraph for resistance of fittings, etc.) The average 
pressure drop in exhaust steam mains varies from 0.2 to 0.4 pound 
per 100 feet for non-condensing service and from 0.2 to 0.4 inch of mer- 
cury per 100 feet for a vacuum of 26 inches. In large steam turbine in- 
stallations there is practically no exhaust piping and steam velocities of 
300-400 feet per second are possible with a negligible pressure drop. 

Notwithstanding the numerous investigations conducted on labora- 
tory apparatus and on pipe lines under actual power plant conditions 
there is no trustworthy rule for accurately determining the behavior of 
the flow of steam in commercial piping. 



742 



STEAM POWER PLANT ENGINEERING 



Table 127 gives some of the rules commonly used in piping design 
and those classified under '^ Group I" have been given particular at- 
tention by various writers. For pressure drops under J lb. per 100 feet 



^^^^i^^^^^;^ 



^' •_ '^^ ''^^^^^^^^^^i^^^^^:^?^^^:^^ 




Fig. 490. General Arrangement of Steam and Exhaust Piping, La Salle 
Hotel, Chicago, III. 

of pipe any of the equations in this group will give results which agree 
fairly well with practice but for greater pressure drops they may lead 
to serious error unless modified to suit the new conditions. 



PIPING AND PIPE FITTINGS 



743 



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►^ 




•^ 




»q 


3 


■^ 


:> 


~~^ 


> 


"^ 


> 


5 



!N 



CO 



^^ 



aw 



•II daOHQ 



'I anoHO 



744 STEAM POWER PLANT ENGINEERING 

It has been shown* that all of the rules in Table 127 with the ex- 
ception of ^'Ledoux" are based on the general equation 

V = C'^ (274) 

and differ only with respect to the assumed value of the coefficient of 
friction. 

In equation (274), 

y = pressure drop, lb. per sq. in., 

C = a coefficient involving a number of reduction constants and the 
coefficient of frictional resistance, 

V = velocity, ft. per second, 

y = mean density of the steam, lb. per cu. ft., 

L = length of straight pipe, ft., or its equivalent, 

d = inside diameter of the pipe, in. 

Equation (274) may be reduced to the form 



w = k\/^, (275) 



L 

in which 

w = weight of steam flowing, lb. per sec, 

k = Si coefficient involving the various reduction constants and the 
coefficients of frictional resistance. 

Numerous experiments have been made with a view of determining 
the coefficient of frictional resistance but the results have been far from 
harmonious. The coefficients involved in the equation given in Table 
127 are not applicable to the present practice of high velocities, pres- 
sures and temperatures. 

Fritzsche's equation (Mitt, uber Forschungsarbeit, Vol. 60) has been 
mentioned as giving results more in accord with current practice but 
recent investigations made at the Berfiner Elektrizitats Werke (Prac. 
Engr., March 15, 1916, p. 284) show that pressure drops calculated by 
means of this equation may be 50 per cent too low. In the light of 
the best evidence available at the present time preference should be 
given to the coefficient as determined by J. M. Spitzglass (Armour 
Engineer, May, 1917). Using the values of the coefficient of friction 
as determined by Spitzglass equation (275) may be reduced to the 
convenient form 



\/^' 



w = K\/^, (276) 

in which X is a coefficient with values as given in Table 128. 
* See Author's paper, Power, June, 1907, p. 377. 



PIPING AND PIPE FITTINGS 



74i 



The author has apphed equation (276) to a number of cases in which 
pressure drops have been determined experimentally and the calculated 
values checked substantially with the test results. The values of K 
given in the table allow a sufficient factor of safety for all fittings which 
do not abruptly change the direction of flow or reduce the pressure 
by throttling. Attempts to include factors for condensation losses 
merely complicate the equation without adding to its accuracy. All 
equations thus far established relative to the flow of steam in pipes are but 
approximations at the best and should be used accordingly. 

Example 70. Determine the diameter of pipe suitable for a 30,000- 
kw. steam turbine lead with operating conditions as follows: Initial 
absolute pressure 265 lb. per sq. in., superheat 200 deg. fahr., length 
of pipe 100 ft., maximum pressure drop in the pipe alone to approxi- 
mate 3 lb. per sq. in. when delivering 330,000 lb. steam per hour. 



w = 



330,000 
3600 



= 91.66, L = 100, p = 3. 



For small pressure drops the density may be assumed as that cor- 
responding to initial pressure, thus y = 0.425 for pi = 265 lb. abs. 
and 200 deg. fahr. superheat. 

Substituting these values in equation (276) 



91.66 = K 



3 X 0.425 



100 



from which i^ = 811 + . 

From Table 128 it will be seen that K for a 14-inch pipe is 800. A 
14-inch pipe would therefore be the nearest commercial size which will 
fulfill the required conditions. 



TABLE 128. 
VALUES OF K FOR VARIOUS PIPE SIZES. 



Nominal Pipe 


R- 


Nominal Pipe 


K. 


Diameter, In. 




Diameter, In. 


1.0 


0.75 


5.0 


60 


1.5 


2.5 


6.0 


97.0 


2.0 


5.1 


8.0 


195.0 


2.5 


8.5 


10.0 


350.0 


3.0 


15.5 


12.0 


550.0 


3.5 


23.0 


14.0 


800.0 


4 


32.5 

1 


16.0 


1100.0 



349. Friction througti Valves and Fittings. — Equations 275 to 276 
and those outlined in Table 127 are strictly applicable only to well- 
lagged pipes, free from sharp bends or obstructions such as valves or 
fittings, which greatly increase the resistance of the flow of steam. If 
these obstructions must be considered, it is customary to allow for 



746 



STEAM POWER PLANT ENGINEERING 



them by assuming an added length of straight pipe equivalent in re- 
sistance to the various fittings and bends. Unfortunately, the few- 
tests which have been made for the purpose of determining the resist- 
ance of various pipe fittings give discordant results, and rules based 
on these investigations are limited to such a narrow range of operating 
conditions that their use for general design purposes is apt to lead to 
serious error. 

It is definitely known that the value of the coefficient of resistance 
for smooth piping decreases with increasing diameter but with globe 
valves in short lengths of piping it appears to increase with increasing 





1 M 1 1 1 1 M 1 1 
























— 


A Total drop-boiler drum to steam lead. 
B 6-in. Automatic Check Valve 
C Superheater 
D Stop Valve 






















/ 




















A 




















/ 


/ 












































/ 


/ 












































/ 


/ 














In 


tia 


Pr 


;ssu 


re 2 


5 1b 


s. a 


3S. 
















/ 


/ 
















Sd 


per 


lea 


160 


dej 


•. f^hr. 
















y\ 












































^ 


y 














^ 




























^ 


^ 










































^ 








B 




^ 


-^ 


























, 


^ 


-^ 








^ 


-^ 


— 


c 


^ 


^ 




















— ' 


' 


_^^^ 











■^ 






, . 


. 


— 


"^ 






D 








. 

















-1 


— ■ 


— 


__ 


__ 










-— 


— 


— 


— ■ 

















1080 



10 15 20 25 

Steam Flow • Thousands of Pounds per Hour 
2110 3260 4370 5510 

Steam Velocity - f^eet per Minute 



Fig. 491. Steam Pressure Drop, 500-hp. Babcock and Wilcox Boilers. 



diameters. It seems probable that the placing of a globe valve in a 
hmited length of piping produces an increasing diminution of the free 
steam passage for increasing valve diameters and thereby causes a 
whirling and friction which increases the resistance. The frequently 
observed fact that piping of large diameter always gives a higher 
pressure drop than is expected from calculations is probably due to 
allowing too small values for the resistance of valves, superheaters 
and fittings. 

According to Briggs ("Warming Buildings by Steam") the length 
of straight pipe in inches equivalent to the resistance of one standard 
90-degree elbow is 

L = 75d--(n-^j, (277) 

and that of a globe valve 

£, = 114d^.(n-MJ. (278) 



PIPING AND PIPE FITTINGS 



747 



These rules have been frequently quoted but results calculated from 
them are not in accord with the actual pressure drop in modern power 
plant practice. The curves in Figs.* 491 and 492 give some idea of the 
pressure drops in the piping system of a modern turbo-generator plant 
and serve to show that the actual drops are much higher than ordi- 
narily supposed. 

350. Equation of Pipes. — It is frequently desirable to know what 
number of one sized pipes will be equal in capacity to another pipe. 

According to the equations in Group II, Table 127, the weights 
discharged for a given set of conditions vary with the square root of 









- 
















ab Boiler dry pipe 
























- 


220 




a 


















cd Stop and Automatic Check Valve 

de 2-7-in. leads and header to hydraulically 

operated valve 
ef Separator and lead to turbine 
fg Throttle Valve 












\ 
























215 






b- 


^ 




^ 














































210 














































I 
























d 








— 


— 




— 


— 







e 




















205 










































^ 


— 






^^ 




/ 








4-1225 


H-p. B. & W 


Boilers \i'ith 


2-'? in 


leads 






















■^ 




200 






18^in. p.I^. s'^eaJD ipea|der 
































\ 








18 000 Kw. Load on Turbine 
































1 


Q 


195 











































































































20 10 GO 80 100 120 140 160 180 200 220 240 260 
Lineal Distance In Feet. 

Fig. 492. Steam Pressure Drop from Boiler Drums to Turbine Throttle. 



the fifth power of the diameter; that is, the number of pipes equal in 
capacity to any given pipe may be determined from the equation 

AT, = ^1 ^ rf,f, (279) 

in which A^i = number of pipes of diameter di equal in capacity to a 
pipe of diameter d\ di and d in inches. 
According to the equation (276) 

r/-' d,' \\ (280) 



"i.tf^^) 



Thus, one 8-inch pipe is equal in capacity to six 4-inch pipes, or 
(from Table 128) A^i = 195 ^ 32.5 = 6. 

351. Exliaust Piping, Condensing Plants. — The exhaust piping in 
condensing plants is arranged cither according to (1) the independent 
or (2) the central condensing system. In the former each engine is 
provided with an independent condenser and air pump. In case the 

* Courtesy of A. D. Bailey, Engineer in Charge of Fisk Street and Quarry Street 
Stations, Commonwealth Edison Co., Chicago. 



748 



STEAM POWER PLANT ENGINEERING 




PIPING AND PIPE FITTINGS 



749 



vacuum ''drops" or it is desired to operate non-condensing, the steam 
is discharged through a branch pipe with reUef valve to the atmosphere, 
Figs. 3 and 321. When there are a number of engines in one installa- 
tion the atmospheric pipes lead to a common free exhaust main, which, 
on account of its great size, is ordinarily constructed of light-weight 
riveted steel pipe. The short connection between engine and condenser 
is usually made with lap-welded steel pipe, since riveted joints are apt 
to leak, due to the engine vibrations. In a central condensing plant, 
Fig. 330, the several engines exhaust through a common main into a 
single large condenser. An atmospheric relief valve is usually provided 
in connection with the condenser, and no free exhaust main is necessary. 
Several arrangements of condenser piping are illustrated in Figs. 321 
to 330. 

352. Exhaust Piping, Nan-condensing Plant. Webster Vacuum System. 
— In the majority of non-condensing plants the exhaust steam is 
used for heating purposes. One of the best-known systems of exhaust 
steam heating, in which the back pressure on the engine is reduced by 
circulating below atmospheric pressure, 
is that known as the Webster combina- 
tion system. The general arrangement 
is illustrated in Fig. 2 and the prin- 
ciples of operation are described in 
paragraph 3. It has the advantage of 
affording (1) minimum back pressure 
on the engine; (2) effective and con- 
tinuous drainage of condensation from 
supply pipes and radiators; (3) con- 
tinuous removal of air and entrained 
moisture from confined spaces; (4) in- 
dependent regulation of temperature in 
each radiator; (5) /continuous return of condensation to the boiler; 
(6) utihzation of part of the exhaust steam for preheating the feed 
water; and (7) automatic regulation. Fig. 493 gives a diagrammatic 
arrangement of the piping and appurtenances in a typical installation. 
The characteristic feature of this system is the automatic outlet valve 
attached to each part requiring drainage, which permits both the water 
of condensation and the non -condensable gases to be removed continu- 
ously. The radiator temperature may be regulated by varying the 
quantity _of steam supplied, either by hand or automatically by thermo- 
static control. The Webster valve. Fig. 493a, enables the vacuum to 
withdraw the water of condensation as fast as it is formed irrespective 
of the pressure in the radiator; hence the supply may l)e throttled to 




OUTLET 

Fig. 493a. Webster Air Valve. 



750 



STEAM POWER PLANT ENGINEERING 




PIPING AND PIPE FITTINGS 



751 



such an extent that the temperature in the radiator is practically as low 
as that of steam corresponding to the pressure in the vacuum line. 
The small annular space between the inner tube of the float F and the 
guide H permits of a vacuum in the body of the valve. When the water 
from the radiator lifts the float the water is drawn into the returns pipe. 
The valve then returns to its seat and the escape of steam is prevented, 
except such as finds its way through the annular space around the guide 
stem H. 

Automatic air valves are constructed in a variety of designs but 
space limitation prevents their description in this work. For a de- 
tailed description of a number of well-known devices consult "Me- 
chanical Equipment of Buildings," by Harding and Willard, John Wile}^ 
& Sons, 1916. 

353. Exhaust Piping, Non-condensing Plants. Paul Heating System. — 
The Paul vacuum system differs from the Webster in that the conden- 
sation, and the air and non- 

' * ^ SUCTION 

condensable gases are sepa- 
rately handled. Referring 
to Fig. 494, which gives a 
diagrammatic arrangement 
of the piping, the condensed 
steam gravitates to the 
automatic returns tank and 
pump and is pumped either 
directly to the boiler or 
through the heater to the 
boiler. Air and vapor arc 
withdrawn from the upper 
part of the radiator by the 
Paul exhauster or ejector E, 
and discharged into the re- 
turns tank, which is vented 
to the atmosphere for the 
escape of the non-condensa- 
l)le gases. The exhauster 
receives its supply of steam 

through pipe 0, Fig. 495, which shows the general arrangement of 
this apparatus. The piping is in duplicate to guard against failure to 
operate. The suction side of the exhauster is connected with the 
air pipes A, A^ Fig. 494. Fig. 496 gives a section through the Paul 
air or vacuum valve which prevents steam from blowing into the air 
pipes and permits only air to pass. In Fig. 494 the heating system is 




STEAM SUPPLY 



Fig. 495. Paul Exhauster. 



752 



STEAM POWER PLANT ENGINEERING 



COMPOSITION 




FROM RADIATOR 



TO EXHAUSTER 

Fig. 496. Paul Vacuum Valve. 



piped on what is known as the *^ one-pipe down-feed" principle; i.e., 
the exhaust steam is first conducted to a distributing header in the attic, 
from which the various supply pipes are led to the radiators. The 
water of condensation returns through these same pipes and gravitates to 

the returns pump. Both the 
supply steam and the condensa- 
tion flow in the same direction. 
This system is also piped on the 
"one-pipe up-feed," the "two- 
pipe up-feed," and the "two- 
pipe down-feed" principle. The 
"one-pipe up-feed" differs from 
the system just described in 
that the steam flows upward 
through the risers and does 
away with the attic piping. The returns, however, flow against the cur- 
rent of steam, and water hammer is more Hkely to occur than with the 
down-feed system. In the two-pipe systems the steam supply pipes or 
risers conduct steam only, and the returns carry the condensation. 
The one-pipe down-feed is cheaper and simpler and practically as efficient 
as the two-pipe system under normal conditions. It is objectionable, 
however, due to the difficulty of draining the radiator with closely 
throttled supply valve, since the velocity of the entering steam prevents 
the water from returning through the same orifice. 

554. Automatic Temperature Control. — Experience shows that a 
considerable saving in fuel may be effected in the heating plants of 
tall office buildings and similar plants by automatically controlling the 
temperature. Hand-controlled valves are usually left wide open, and 
when the room becomes too hot the temperature is frequently lowered 
by opening the window, resulting in a waste of heat which may be 
considerable in modern buildings with hundreds of offices. Many 
successful methods of automatic temperature control are available, the 
usual system consisting of thermostats which control the supply of heat 
by means of diaphragm valves, the latter taking the place of the usual 
radiator supply valve. 

Fig. 497 shows a Powers thermostat. The expansible disk U contains 
a volatile liquid having a boiling point of about 50 deg. fahr. The 
pressure of the vapor within the disk at a temperature of 70 degrees 
amounts to six pounds to the square inch, and varies with every change 
of temperature, causing a variation in the thickness of the disk. The 
disk is attached by a single screw to the lever Q, which rests upon 
the screw F as a fulcrum. The flat spring R holds the lever and disk 



PIPING AND PIPE FITTINGS 



753 



.^sss 



ssssssssssssss 



^W^Wnn^^ 




against the movable flange M. Connecting with the chamber A^ are 
two air passages H and /. The thermostat is attached by means of two 
screws at the upper end to a wall plate permanently secured to the wall. 
This wall plate has ports registering with H and /, one for supplying 
air under pressure and the other for conduct- ^ 
ing it to the diaphragm motor w^hich operates 
the valve or damper. Air is admitted through 
H under a pressure of about fifteen pounds '"' 
per square inch, and its passage into cham- 
ber N is regulated by the valve J, which is 
normally held to its seat by a coil spring 
under cap P. K is an elastic diaphragm car- 
rying the flange M, with escape valve passage 
covered by the point of valve L. Valve L 
tends to remain open by reason of the spring. ^ 
When the temperature rises sufficiently ex- 
pansion of the disk U first causes the valve 
to seat, its spring being weaker than that 
above valve J. If the expansive motion is 
continued, valve J is lifted from its seat and 
compressed air flows into chamber A^, ex- 
erting a pressure 
upon the elastic 
diaphragm K in 
opposition to the 
expansive force 

of the disk If ^^^- ^^^- Section through 
the temperature ^"^^'^ Thermostat. 

falls, the disk contracts and the over- 
balancing air pressure in N results in a 
reverse movement of the flange M, per- 
mitting the escape valve to open and dis- 
charge a portion of the air; thus the air 
pressure is maintained always in direct 
proportion to the expansive power (and 
temperature) of the disk U. The passage 
I communicates with a diaphragm valve, 
Fig. 498. The compressed air operates the 
diaphragm against a coiled spring resistance, so that the movement 
is proportional to the air pressure and the supply of steam controlled 
accordingly. The adjusting screw G, squared to receive a key, carries 
an indicator by means of which the thermostat can be set to carry any 




Fig. 498. 



A Typical Diaphragm 
Valve. 



754 



STEAM POWER PLANT ENGINEERING 



desired temperature within its range, usually from 60 to 80 degrees. 
In changing the temperature adjustment lever Q forces the disk U 
closer to or farther from the flange M. 

In connecting up the system compressed air is carried to the thermo- 
stat and diaphragm valves, from a reservoir through small concealed 
pipes. 

In the indirect system of heating the dampers are of the diaphragm 
type and the method of regulation is the same as with the direct system. 

355. Feed-water Piping. — The simplest arrangement of feed-water 
piping may be found in non-condensing plants, in which the feed water 
is obtained under a slight head, such as is afforded by the average city 
supply, and is heated in an open heater by the exhaust steam from the 
engine to a temperature varying from 180 to 210 deg. fahr. The hot 



OQ 



\:: 



tt 



ft 




o^ 



INJECTOR MAIN 



Uio 



•..__-__5. 



• j ' I ' I r eeo PUMP I I rEE( 



1 Fteeo pump 




COLD WATtH SUPPty 



-t 



Fig. 499. Feed-water Piping; Non-condensing Plant. 

feed water gravitates from the heater to the pump and then is forced 
to the boiler, or to the economizer if one is used. If a meter is used 
it is generally placed on the discharge side of the pump, and should be 
by-passed to permit it to be cut out for repairs (Fig. 499). Plants 
operating continuously should have feed pumps in duplicate. In some 
cases the returns from the heating system gravitate to the heater and 
only enough cold water is added to make up the loss from leakage, etc. ' 
In other cases the returns gravitate to a special '' returns tank," from 
which they are pumped directly to the boiler without further heating. ! 
Occasionally a live-steam purifier is used, especially if the water contains 
a large percentage of calcium sulphate. The feed is then subjected to 
boiler pressure and temperature and the greater part of the impurity | 
precipitated before it enters the boiler. Closed heaters are often used 
in place of open heaters. When the supply is not under head a closed 
heater is usually preferred and is placed between the pump discharge 
and the feed main. 



PIPING AND PIPE FITTINGS 



755 



In condensing plants the feed piping is similar to that in non-con- 
densing plants, except that if exhaust steam is used for heating purposes 
it is supplied by the auxiliaries, such as feed pumps, stoker engines, 
condenser engines, and other steam-using appliances. 

In plants having a number of boilers it is customary to run a feed 
main or header the full length of the boiler room and connect it to each 
boiler by a branch pipe. This main may be a simple header or in 
dupHcate or of the ''loop" or "ring" type. Horizontal tubular boilers 
are frequently arranged in one battery with the feed main run along the 
fronts of the boilers just above the fire doors. Water-tube boilers are 
generally set in a battery, and as the arrangement above would block 
the passageway between the batteries, the main is run either above or 
under the settings, the former being the more common. Where a 



QQ O© 



L 







Fig. 509. Feed-water Piping; Condensing Plant. 

single header is used, the feed pumps are sometimes placed so as to 
feed into opposite ends of the main, which is then cut into sections by 
valves. Another arrangement is to place the pumps so as to feed into 
the middle of the header. With the loop arrangement the main is 
ordinarily cut into sections by valves so that the water may be sent 
either way from the pumps and any defective section cut out. With 
duplicate mains a common arrangement is to place one main along the 
front of the boiler and the other at the rear or both overhead as in 
Fig. 489. Sometimes one main is placed in the passageway below the 
boiler setting and the other on top. 

Standard wrought-iron pipe is usually used for pressures under 100 
pounds and extra heay>^ pipe for greater pressures. The pipes and 
fittings from boiler to main are frequently of brass, and preferably so, 
since brass withstands corrosive action much better than iron or steel. 
Flanged joints should be used in all cases, since the pockets formed by 
the ordinar^^ screwed joints hasten corrosion at those points. (Power, 
June, 1902, p. 4.) 



756 



STEAM POWER PLANT ENGINEERING 



Fig. 502, A to E, illustrates the various combinations of check valve, 
stop valves, and regulating valve in steam boiler practice. The simplest 
arrangement and one sometimes used in plants operating intermittently 




fSZZZS7;^7^77777777ZZZ^7ZZ^^77Z^Z7Wy 



Fig. 501. Feed-water Piping, 



is shown in A. Here there are but two valves between the boiler and 
the main, the check being nearest the boiler and the stop valve at the 
main. The stop valve performs both the function of cutting out the 




Main Feed Header 

Fig. 502. Different Arrangements of Valves in Feed- water Branch Pipes. 

boiler and of regulating the water supply. This arrangement is not 
recommended, as any sticking or excessive leaking of the check valve 
will necessitate shutting down the boiler. B shows the most common 
arrangement. Here the check valve is placed between the regulating 



i 



PIPING AND PIPE FITTINGS 757 

valve and a stop valve as indicated. This permits a disabled check to 
be easily removed while pressure is on the boiler and the main. E shows 
an arrangement whereby both check and regulating valve may be 
removed, and is particularly adapted to boilers operating continuously 
where the regulating valve is subjected to severe usage. In this case 
the stop valves are run wide open and are subjected to no wear. The 
regulating valve most highly recommended is a self-packing brass globe 
valve with regrinding disk. The check valve is ordinarily of the swing 
check pattern with regrinding disk, Fig. 513 (C). Modern practice 
recommends an automatic water relief valve in the discharge pipe im- 
mediately adjacent to each pump (piston type only) to prevent excessive 
pressure in case a valve is accidentally closed in by-passing or in changing 
over. 

356. Flow of Water through Orifices, Nozzles, and Pipes. — Bernoulli's 
theorem is the rational basis of most empirical equations for the steady 
flow of a fluid from an up-stream position n to a down-stream position m, 
thus (''Mechanics of Engineering," Church, p. 706): 



P y 2 p yn2 

•* wi 1 ' m 1 r7 ■'- n I ' 1 '7 

y 2g 7 2g 



all losses of head 
occurring between 
n and m 



(281) 



in which 



V = velocity in feet per second at the point considered, 

P = pressure in pounds per square foot, 

Z = potential head in feet of the fluid, 

7 = density of the fluid, pounds per cubic foot, 

g = acceleration of gravity. 

y2 
Each loss of head wiU be of the form K -^r- m which K is the coefficient 

of resistance to be determined experimentally. The loss of head due to 
skin friction is expressed : 

H = 4/^x|l, (282) 

in which 

/ = the coefficient of friction of the fluid in the pipe, 
I = length of the pipe in feet, 
d = diameter of the pipe in feet. 

Other notations as in (281). 

Discharge from a circular vertical orifice with sharp corners: 

Q = CA V2^h, (283) 



758 STEAM POWER PLANT ENGINEERING 

in which 

Q = cubic feet per second, 

C = coefficient, varying from 0.59 to 0.65 (Merriman, "Treatise on 

HydrauHcs," p. 118), 
A = area of the orifice, square feet, 
h = head of water in feet, 
g = acceleration of gravity = 32.2. 

Discharge from short cylindrical nozzles three diameters in length, with 
rounded entrance (''Mechanics of Engineering," Church, p. 690): 

Q = 0M5 A V2^h. (284) 

Discharge from short nozzles with well-rounded corners and conical 
convergent tubes, angle of convergence 13^ degrees (Church, p. 693) : 

Q = 0.94 A V2^. (285) 

Discharge from cylindrical pipe under 500 diameters in length (Church, 

p. 712): 

d^ 
(1 + 0.5) d + 4/Z 
in which 



Q = 6.3 y ^ , n.V:. ■ .r7 > (286) 



/ = coefficient of friction. 

Other notations as above. 

/ varies with the nature of the inside surface, the diameter of the 
pipe, and the velocity of flow. 

Discharge through very long cylindrical pipes (''Mechanics of Engineer- 
ing," Church, p. 715): 

Q = 3.15\/^. (287) 

Loss of head due to friction in water pipes* Weisbach^s equation is 
as follows: 

in which 



H = (o.( 



H = friction head in feet, 

V = velocity in feet per second, 

L = length of pipe in feet, 

d = diameter of pipe in inche: 

* See also, Friction Formulas for Commercial Pipe, by Ira N. Evans, Power, 
July 9, 1912, p. 54. 



PIPING AND PIPE FITTINGS 



759 



TABLE 129. 

TABLE OF THE COEFFICIENT / FOR FRICTION OF WATER IN CLEAN 

IRON PIPES. 

(Abridged from Fanning.) 





Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Velocity in 
Ft. per Sec. 


= iin. 

= .0417 ft. 


= 1 in. 
= .0834 ft. 


= 2 in. 
= .1667 ft. 


= 3 in. 

= .25 ft. 


= 4 in. 
= .333 ft. 


= 6 in. 
= .50 ft. 


= 8 in. 
= .667 ft. 


0.1 


.0150 


.0119 


.00870 


.00800 


.00763 


.00730 


.00704 


0.3 


.0137 


.0113 


850 


784 


750 


720 


693 


0.6 


.0124 


.0104 


822 


767 


732 


702 


677 


1.0 


.0110 


.00950 


790 


743 


712 


684 


659 


1.5 


.00959 


.00868 


.00757 


.00720 


.00693 


.00662 


.00640 


2.0 


.00862 


810 


731 


700 


678 


648 


624 


2.5 


795 


768 


710 


683 


662 


634 


611 


3.0 


.00753 


.00734 


.00692 


.00670 


.00650 


.00623 


.00600 


4.0 


722 


702 


671 


651 


631 


607 


586 


6.0 


689 


670 


640 


622 


605 


582 


562 


8.0 


663 


646 


618 


600 


587 


562 


544 


12.0 


630 


614 


590 


582 


560 


540 


522 


16.0 


.00618 


.00600 


.00581 


.00570 


.00552 


.00530 


.00513 


20.0 


615 


598 


579 


566 


549 


525 


508 



Velocity in 
Ft. per Sec. 


Diam. 


Diam. 


Diam. 


Diam. 


riiam. 


Diam. 


Diam. 


= 10 in. 


-=12 in. 


= IG in. 


= 20 in. 


= 30 in. 


= 40 in. 


= 60 in. 


= .833 ft. 


= 1.00 ft. 


= 1.333 ft. 


= 1.667 ft. 


= 2.50 ft. 


= 3.333 ft. 


= 5. ft. 


1 


.00684 
673 
659 


.00669 
657 
642 


.00623 
614 
603 










3 


.00578 
567 








0.6 


.00504 


.00434 


.00357 


1.0 


643 


624 


588 


555 


492 


428 


353 


1.5 


.00625 


.00607 


.00572 


.00542 


.00482 


.00421 


.00349 


2.0 


609 


593 


559 


529 


470 


416 


346 


2.5 


596 


581 


548 


518 


460 


410 


342 


3 


.00584 


.00570 


.00538 


.00509 


.00452 


.00407 


.00339 


4.0 


568 


553 


524 


498 


441 


400 


333 


6.0 


548 


534 


507 


482 


430 


391 


324 


8.0 


532 


520 


491 


470 


422 


384 


320 


12.0 


512 


500 


478 


457 


412 


377 


.00313 


16.0 


.00502 


.00491 


.00470 


.00450 


.00406 


.00370 




20.0 


498 


485 

























William Cox (American MacTiinist, Dec. 28, 1893) gives a simple 
rule which gives almost identical results : 

(4 y2 + 5 7 _ 2) L 



H = 



(289) 



1200 d 
Notations as in (288). 

Loss of head due to friction of fittings. Equations (286) to (289) are 
based on the flow of water through clean straight cylindrical pipes. 
Where there are bends, valves, or fittings in the line the flow is de- 
creased on account of the additional resistance. 



760 STEAM POWP]R PLANT ENGINEERING 

These frictional losses are conveniently expressed in feet of water, 

thus: y2 

H-C^^, (290) 

C having the following values: 

Angles. Class of Valve. 



45 degrees. 90 degrees. Gate. Globe. Angle. 

C 0.182 0.98 0.182 1.91 2.94 

Example 71. Determine the pressure necessary to deliver 200 gal- 
lons of water per minute through a 4-inch iron pipe line 400 feet long, 
fitted with four right-angle elbows and two globe valves. The water 
is to be discharged into an open tank. 

A flow of 200 gallons per minute gives a velocity of 

—75 X7C 777^ = 5 feet per second (7.48 = number of gallons per 

7.4o X bU X IZ.iZ 

cubic foot, and 12.72 = internal area of the pipe, square inches). 

From the preceding table, / = 0.00618 for V = 5. 

From (290), 

25 

Resistance head of 4 elbows = 0.98 X jrr-: X 4 = 1.52 feet. 

64.4 

Resistance head of 2 globe valves: 

1-91 X p;^ X 2 = 1.48 feet. 
64.4 

Resistance head of all fittings: 

1.52 + 1.48 = 3 feet. 
Substitute 7 = 5, L = 400, and ^ = 4 in (289). 

^4 X 52 + 5x5-2 



"=(■ 



400 



1200 X 4 
= 10.25 feet, resistance head of the pipe. 

Total resistance head = 10.25 + 3 = 13.25 feet of water, or 5.75 
pounds per square inch. 

Example 72. How many gallons of water will be discharged per 
minute through above line with initial pressure of 100 pounds per square 
inch, and what will be the pressure at the discharge end? 

Since / depends upon the unknown V, we may put / = 0.006 for a 
first approximation and solve for V; then take a new value of / and 
substitute again, and so on. 

Substitute / = 0.006, d = j\, h = 100 X 2.3 = 230, and I = 400 in 
(287): 

Q 



- "= s/ol 



33^ X 230 



006 X 400 

1.95 cubic feet per second, corresponding to 
a velocity of 22 feet per second. 



PIPING AND PIPE FITTINGS 



761 



From table, 

/ = 0.00548 (by interpolation) for V = 22 feet per second. 

From (290) the friction of 4 elbows and 2 globe valves is found to 
be 58 feet for V = 22. 

From (289) a resistance head of 58 feet of water for F = 22 is found 
to be equivalent to 136 feet of straight pipe, thus: 

X 222 X 5 X 22 - 2\ . 



58 



=(- 



1200 X 4 



L = 136. 
Substitute / = 0.0548, ^ = 400 + 136 = 536 in (287) : 



Q = 3.15V/^ 



33^ X 230 



,0058 X 536 
= 1.74 cubic feet per second, corresponding to 

a velocity of 19.3 feet per second. 
= 780 gallons per minute. 

If greater accuracy is necessary determine / and L for V = 19.3 and 
proceed as above. 

The total friction head may be determined from (289), thus: 

X 19.32 + 5 X 19.3 - 2> 



•"{' 



1200 X 4 
= 177 feet of water 
= 77 pounds per square inch. 



536 



The pressure at the discharge end will be 

100 — 77 = 23 pounds per square inch. 

Average power plant practice gives the following maximum velocities 
of flow in water pipes: 



Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


Jtoi 
itoli 
li to 3 


50 
100 
200 


3 to 6 
Over 6 


250 
300-400 



357. Stop Valves. — The valves used to control and regulate the 
flow of fluids are the most important element in any piping system. 
A good valve should have sufficient weight of metal to prevent distor- 
tion under varying temperature and pressure, or under strains due to 
connection with the piping; the seats should be easily repaired or re- 
newed; there should be no pockets or projections for the accumulation 
of dirt and scale, and the valve stem should permit of easy and efficient 



762 



STEAM POWER PLANT ENGINEERING 



packing. Stop valves are made in such a variety of designs that a 
brief description will be given of only a few fundamental types. 

Fig. 503 shows a section of an ordinary globe valve, so called because 
of the globular form of the casing. This type of valve is the most 
common in use. Globe valves are designated as (1) inside screw and 
(2) outside screw, according as the screw portion of the stem is inside 
the casting, Fig. 503, or outside, Fig. 504. The top, or bonnet, may be 
screwed into the body of the valve. Fig. 503, or bolted, Fig. 504. The 
smaller sizes, three inches and under, are usually of the screw-top type 
and the larger of the holt-top type. Valves with outside yoke and screw 





Fig. 503. A Typical Globe Valve, 
Screw-top, Inside Screw. 



Fig. 



504. A Typical Globe Valve, 
Bolt-top, Outside Screw. 



are preferable to others in that they show at a glance whether the 
valve is open or closed, an advantage in changing from one section to 
another. The disks are made in a variety of forms, the material 
depending upon the nature of the fluid to be controlled. Thus, for 
cold water, hard rubber composition gives good results; for hot water 
and low-pressure steam. Babbitt metal; for high-pressure steam, copper 
or bronze ; and for highly superheated steam, nickel. The valve bodies 
are of brass for sizes under three inches, cast iron for the larger sizes 
and ordinary pressures and temperatures, and cast steel or semi-steel 
for high temperatures and pressures. Globe valves should always be 
set to close against the pressure, otherwise they could not be opened 
if the valves should become detached from the stem. Globe valves 



PIPING AND PIPE FITTINGS 



763 



should never be placed in a horizontal steam return pipe with the stem 
vertical, because the condensation will fill the pipe about half full be- 
fore it can flow through the valve. Globe valves that are open all the 
time are preferably designed with a self-packing spindle, as in Fig. 504, 
in which the top of shoulder C can be drawn tightly against the under 
surface of bonnet S, thus preventing steam from leaking past the screw 
threads while the spindle is being packed. 

Figs. 505 to 507 show different types of gate or straightway valves. 
These valves offer little resistance to the flow of steam or liquid passing 




Fig. 505. A Typical 
Gate Valve, Solid- 
wedge, Screw-top, 
Outside Screw. 



Fig. 506. A Typical Gate 
Valve, Solid-wedge, Bolt- 
top, Inside Screw. 



Fig. 507. A Typical 
Gate Valve, Split- 
wedge, Bolt-top, In- 
side Screw. 



through them, and are generally used in the best class of work. Fig. 
505 shows a section through a solid-wedge gate valve with outside screw 
and yoke. This form of outside screw and yoke with stem protruding 
beyond the hand wheel is a perfect indicator to show whether the valve 
is open or shut, as the hand wheel is stationary and the spindle rises in 
direct proportion to the amount the valve is opened. For these reasons 
outside screw valves are preferable for high-pressure work and especially 
for the larger sizes. The seats are made solid, or removable, and of 
various materials for different pressures and temperatures. Fig. 507 
shows a section through a split-wedge gate valve with parallel faces and 
seats. For the sake of illustration this valve is fitted with inside screw. 



764 



STEAM POWER PLANT ENGINEERING 



In this design the spindle remains stationary so far as any vertical 
movement is concerned, and the gate or plug, being attached to it by 
means of a threaded nut, rises into the bonnet when the spindle is re- 
volved. It is impossible to tell by its appearance whether this form of 
valve is opened or closed. Valves with inside screw are adapted to 
situations where there is considerable dirt and grit, since the screw is 
inclosed and protected, and excessive wear is thus avoided. Gate 
valves with split gates are more flexible than those with solid gates, and 
hence are less likely to leak. Fig. 508 shows the application of the 





Fig. 508. Ludlow Angle Valve, 
Gate Pattern. 



Fig. 



509. Anderson Non- 
return Valve. 



gate system to an angle valve. All high-pressure valves above 8 inches 
in diameter should be provided with a small by-pass valve, as the 
pressure exerted against the disk or gate is very great when the valve is 
closed and the force required to move it is considerable. The by-pass 
valve also facihtates ''warming up" the section to be cut in and is 
more readily operated than the main valve. 

358. Automatic Non-return Valves. — Fig. 509 shows a section 
through an automatic non-return valve as applied to the nozzle of a 
steam boiler. As will be seen from the illustration it practically 
amounts to a large check valve with cushioned disk. The object of 
this device is the equalization of pressure between the different units of 



PIPING AND PIPE FITTINGS 765 

the battery, the valve remaining closed as long as the individual boiler 
pressure is lower than that of the header. In case a tube blows out the 
valve closes automatically, owing to the reduction of pressure, and 
prevents the header steam from entering the boiler. It acts also as a 
safety stop to prevent steam being turned into a cold boiler while men 
are working inside, because it cannot be opened when there is pressure 
on the header side only. To be successful, such a valve should not 
open until the pressure in the boiler is equal to that in the header; it 
should not stick and become inoperative nor chatter and hammer while 
performing its work. Referring to Fig. 509, tail rod E insures align- 
ment and hence prevents sticking; steam space C acts as a dashpot to 
prevent hammering of the valve as it rises, and steam space D acts as 
a cushion and prevents hammering at closing. Lip F is made to enter 
the opening in the seat and reduce wire drawing across the seat. Fig. 
485 shows the installation of a number of non-return valves at the 
Yonkers power house of the New York Central Railway Company. 

359. Emergency Valves and Automatic Stops. — In large power 
plants it is customary to protect the various divisions of the steam 
piping by emergency valves which may be closed by suitable means at 
any reasonable distance from the valve. The simplest form of emer- 
gency stop is a weighted "butterfly" valve, which is to all intents and 
purposes a weighted check, as illustrated in Fig. 513 (D). The weight 
when supported, say by a cord and pulley, holds the valve open; when 
the cord is cut or released the weight drops and forces the valve shut. 
The cord may lead to any convenient and safe distance from the valve. 
In applying this system of control to steam engines the valve is placed 
in the steam pipe just above the throttle and the weight held up by 
a lever controlled by the main governor or preferably by a separate 
governor. Should the engine exceed a certain speed, as in case of 
accident to the regular governor, the lever supporting the weight is 
tripped by the emergency governor and the valve is closed automati- 
cally. For high pressures a rotating plug valve or cock is preferred to 
the butterfly type, since it is balanced in all positions. Gate and globe 
valves may be converted into emergency valves by having the stems 
mechanically operated by electric motors, hydraulic pistons, and the 
like. Fig. 510 shows a section through a Crane hydraulically operated 
emergency gate valve. 

Fig. 511 shows a partial section through an ''Anderson triple-duty" 
emergency valve, and Fig. 512 a section through the pilot valve. A 
steam connection from the main line to the top of a copper diaphragm 
holds the pilot valve closed because of the large area above the dia- 
phragm. A steam pipe connection from underneath the emergency 



766 



STEAM POWER PLANT ENGINEERING 



piston of the triple-acting valve also leads to the pilot valve. In case 
a break occurs in the main steam line or branches, the pressure is re- 
moved from the top of the pilot valve, causing it to open, thus exhaust- 
ing the pressure from beneath the emergency piston in the triple- 
acting valve. The boiler pressure on top of the emergency piston 






Fig. 510. Crane 
Emergency 
Valve, Hydrau- 
lic 



Fig. 511, Anderson 
Triple-duty Emer. 
gency Valve. 



Fig. 512. Pilot Valve 
for Anderson Triple- 
duty Emergency 
Valve. 



causes' the valve to close. Pilot valves may be located at any desirable 
places, thus affording control from different points. 

In the ''Locke automatic engine stop system" the stop valve is 
operated by an electric motor which is controlled by contact points 
operated by a speed-hmit device. (See Power, August, 1907, p. 471, 
for a detailed description.) 

360. Check Valves. — Fig. 513, A to D, illustrates the different types 
of check valves in most common use. A is a hall check, B a cup or disk 
check, C a siving check, and D a weighted check. Occasionally the valve 
body is fitted with a valve stem and handle for holding the disk against 
its seat, in which it is designated as a slop check. In A and B the valve 



PIPING AND PIPE FITTINGS 



767 



scat is parallel to the direction of flow and the valve is held in place 
by its own weight and by the pressure of the fluid in case of reverse flow. 
In the swing check the seat is at an angle of about 45 degrees to the 
direction of flow. The latter construction is preferred as it offers less 
resistance to flow and there is less tendency for impurities to lodge on 
the valve seat. By extending the hinge of the swing through the body 
of the valve, a lever and weight may be attached as in Z) and the check 
will not open except at a pressure corresponding to the resistance of 
the weight. It thus acts as a reUef valve and at the same time pre- 
vents a reversal of flow. Stop checks are usually inserted in boiler feed 
lines close to the boiler, and, when locked, act as any ordinary stop 




(A) 



(B) (c) 

Fig. 513. Types of Check Valves. 



(DJ 



valve and permit the piping to be dismantled or the regulating valve 
to be reground without lowering the pressure on the boiler. Since the 
wear on check valves is excessive and necessitates frequent regrinding 
they are often mounted with regrinding disks, Fig. 513 (C), which may 
be ''ground" against the seat without removing the valve from the line. 

361. Blow-oflf Cocks and Valves. — The requirements of a good blow- 
off valve are that it shall furnish a free passage for scale and sediment, 
that it shall close tightly so as not to leak, and that it shall open easily 
without sticking or cutting. On account of the rather severe service 
to which such valves are subjected, they are made very heavy, with 
renewable wearing parts. 

Fig. 514 gives a sectional view of a Crane ferrosteel valve. The bon- 
net is easily taken off and the disk removed to be refaced or replaced 
by a new one. The old disk is repaired by pouring in a hard Babbitt 
metal and facing it off flush. The seats are of brass and oval on top 
to prevent scale lodging between them and the disk, and are so made 
that they may be removed; but it has been found in practice that there 
is not much cutting of the seat, the damage usually being confined to 
the softer Babbitt metal which faces the disk. 

Fig. 515 gives a sectional view of a Faber valve. When the disk, 
which makes a snug fit in the body of the valve, is in the position shown, 
the boiler discharge is practically shut off and any sediment lying on 
the seat is cleaned off by a jet of steam or water. 



768 



STEAM POWER PLANT ENGINEERING 



Fig. 516 shows a section through a typical hlow-of cock of the straight- 
way taper plug pattern with self-locking cam. Plug cocks are often 
used instead of valves on the blow-ofT piping. 

Current practice recommends the use of two valves, or rather one 
valve and one cock, in the blow-off line of each boiler. In most of the 




Fig. 514. Crane Ferro- Fig. 515. A Typical Blow- 
steel Blow-off Valve. off Cock. 



Fig. 516. Faber Blow- 
off Valve. 



large stations a blow-off valve and a blow-off cock are installed as in- 
dicated in Fig. 517. The number and size of blow-off cocks are usually 
specified by city or state legislation. (For a description of various 
types of blow-off valves, see Power, Dec. 20, 1910, p. 2228.) 

363. Safety Valves. — Fig. 518 shows a section through the simplest 
form of safety valve. The valve is held on its seat against the boiler 
pressure by a cast-iron weight as indicated. This type has the advan- 
tage of great simplicity, and can be least affected by tampering, since 
it requires so much weight that any additional amount which would 
seriously overload it can be quickly detected. For high pressure and 
large sizes of boiler this class of valve is entirely too cumbersome. 

Fig. 519 shows the general details of the common lever safety valve. 
The valve is held against its seat by a loaded lever, thereby enabling 
the use of a much smaller weight than the ^'dead-weight" type, since 
the resistance is multipKed by the ratio of the long arm of the lever 
to the short one. The proper position of the weight is determined by 
simple proportion. Safety valves of the ''dead-weight" or "lever" 
type are httle used in modern practice, and their use is prohibited in 
U. S. marine service and in many states. 



PIPING AND PIPE FITTINGS 



769 




770 



STEAM POWER PLANT ENGINEERING 



Fig. 520 shows a section through a typical pop safety valve in which 

the boiler pressure is resisted by a 
spring. This type of valve has prac- 
tically supplanted all other forms. 
The boiler pressure acting upon the 
under side of valve V is resisted by 
the tension in spring S. As soon as 
the boiler pressure exceeds the resist- 
ance of the spring the valve lifts from 
its seat and the steam escapes through 
opening 0. The static pressure of 

the steam plus the force of its re- 
FiG. 518. "Dead-weight "Safety Valve, ,. • r, • j n j. j r j.t- 

^ ■^ action m bemg deflected from the 

surface A holds the valve open until the pressure in the boiler drops 





Fig. 519. Common Lever Safety Valve. 

about 5 pounds below that at which the valve is lifted, 
tional area of valve exposed to pres- 
sure when the valve lifts causes it to 
open with a sudden motion which has 
given it its name, and it also closes 
suddenly when the pressure has fallen. 
These valves are arranged so that the 
spring tension may be varied with- 
out taking them apart, and provision 
is made for lifting the seats by means 
of a lever. The seats are of solid 
nickel in the best designs, to minimize 
corrosion. 

The commercial rating of a safety 
valve is based upon the area exposed 
to pressure when the valve is closed. 

The number and size of safety valves 



The addi- 



]c§) 



BLOW Orr LIVER 




BOILER CONNECTION 



Fig. 520. Consolidated Pop 
Safety Valve. 



for a given boiler are ordinarily specified by city or state legislation. 



PIPING AND PIPE FITTINGS 771 

The logical method for determining the size of safety valves is to 
make the actual opening at discharge sufficient to take care of all 
steam generated at maximum load without allowing the pressure to 
rise more than six per cent above the maximum allowable working 
pressure, thus: 

Let W = maximum weight of steam discharged, pounds per hour, 

A = effective discharge area, square inches, 

P = boiler pressure, pounds per square inch absolute, 

L = lift of valve, inches, 

K = coefficient determined by experiment, 

D = diameter of valve, inches. 

According to Napier's rule for the discharge of steam through un- 
restricted orifices 

W = ^ PA = 51.4 PA. (291) 

Allowing for restriction of orifice 

W = 51AKPA. (292) 

In the A.S.M.E. ''boiler code" the value of K is taken as 0.96. Sub- 
stituting this value K in equation (292), 

W = 49.3 PA. (293) 

A = ttDL, 
whence W = 155 PDL (294) 

and D = 0.00645 ^- (295) 

For the almost universal 45-degree seated valve 
A = ttDL sine 45 degrees 
= 0.707 DL, 
whence W = 109.7 PDL (296) 

and D = 0.00911 —- (297) 

The present rule of the United States Board of Supervising Inspec- 
tors is 

a = 0.2074^, (298) 

in which 

a = area of the safety valve in square inches per square foot of grate 

surface per hour, 
w = pounds of water evaporated per square foot of grate surface 

per hour. 



For a flat-seated valve 



772 STEAM POWER PLANT ENGINEERING 

Example 73. A boiler at the time of maximum forcing uses 2150 lb. 
of Illinois coal per hour; heat value 12,100 B.t.u. per lb.; boiler pressure 
225 lb. per sq. in. gauge; feed water 200 deg. fahr. Required the size of 
safety valve. 

Assuming a boiler efficiency of 75 per cent the total maximum 
evaporation is 

w 2150 X 12,100 X 0.75 .qqq^,, , 

W = Yc^ = 18,880 lb. per hour. 

(1033 = heat content of 1 lb. of steam at 225 lb. gauge above 200 
deg. fahr.) 

Assuming a lift of 0.1 in., we have, from equation (297), 

According to the A.S.M.E. code two valves would be required. 

Considering two valves of the same size, the diameter of each for the 

7.17 
given condition would be -^— = 3.5 (approx.). 

The following rules pertaining to safety valves are taken from the 
A.S.M.E. Boiler Code: 

Each boiler shall have two or more safety valves, except a boiler 
for which one safety valve 3-in. size or smaller is required. 

One or more safety valves on every boiler shall be set at or below 
the maximum allowable working pressure. The remaining valves may 
be set within a range of three per cent above the maximum allowable 
working pressure, but the range of setting of all of the valves on a 
boiler shall not exceed ten per cent of the highest pressure to which 
any valve is set. 

Each valve shall have full sized direct connection to the boiler. No 
valve of any description shall be placed between the safety valve and 
the boiler, nor on the discharge pipe between the safety valve and 
the atmosphere. 

The complete A.S.M.E. Boiler Code may be purchased from the 
American Society of Mechanical Engineers, New York City. 

363. Back-pressure and Atmospheric Relief Valves. — These valves 
are for the purpose of preventing excessive back pressure in exhaust 
pipes. In non-condensing plants such valves are designated as back- 
pressure valves and in condensing plants as atmospheric relief valves. In 
the former the valve is usually adjusted so that a pressure of one to 
five pounds above the atmosphere is necessary to lift it from its seat; 
in the latter the valve lifts at about atmospheric pressure. They are 
practically identical in construction, differing only in minor details. 
A slight leakage in the back-pressure valve is of small consequence, 
but in an atmospheric relief valve it may seriously affect the degree 
of vacuum and throw unnecessary work upon the air pump, hence it 






PIPING AND PIPE FITTINGS 



773 



is customary to 'Svater-seal" the latter. Fig. 521 shows a section 
through a typical back-pressure valve. The valve proper consists of 
a single disk moving vertically. The valve stem is in the form of a 





FiQ. 521. Foster Back-pressure Valve. Fig. 522. 



INLET 

Davis Back-pressure Valve. 



piston or dashpot which prevents sudden closing or hammering. The 
pressure holding the valve against its seat is regulated by a spring. 
When the back pressure becomes greater than atmospheric plus that 
added by the spring, the valve raises from its seat and relieves it. 



yff;^.%\'7 





Fig. 523. Crane Atmospheric 
Relief Valve. 



Fig. 524, Acton Atmospheric 
Relief Valve. 



Fig. 522 shows a section through a Davis back-pressure valve, in 
which the resisting pressure is varied by means of a lever and weight. 

Fig. 482 shows the application of a back-pressure valve to a typical 
heating system. 



774 



STEAM POWER PLANT ENGINEERING 



Fig. 523 shows a section through a typical atmospheric relief valve. 
Opening B is connected to the exhaust pipe and opening A leads to 
the atmosphere. Under normal conditions of operation atmospheric 
pressure holds valve V against its seat. Water in groove S ''water- 
seals" the seat and prevents air from being drawn into the condenser. 
In case the pressure in pipe B becomes greater than atmospheric it 
lifts valve V from its seat and is reheved. Piston P acts as a dashpot 
and prevents the valve from slamming. 

Fig. 524 shows a section through an atmospheric relief valve in which 
the weight of the valve is counterbalanced or even overbalanced by an 
adjustable weight and lever, thereby permitting the valve to open at 
or below atmospheric pressure, as may be desired. 

364. Reducing Valves. — It is often necessary to provide steam at 
different pressures in the same plant, as in the case of a combined 





Fig. 525. Kieley Reducing Valve. Fig. 526. Foster Pressure Regulator. 



power and heating plant. To effect this result the reduction in pres- 
sure is accomplished by passing the steam through a reducing valve, 
which is but an automatically operated throttle valve. There are 
many different forms, the operation of all being based upon the same 
general principles. 

In the Kieley valve, Fig. 525, the low-pressure steam acts upon the 



PIPING AND PIPE FITTINGS 775 

top of flexible diaphragm D, and the weighted lever L (which may be 
adjusted to give the desired reduction in pressure) acts upon the other 
side. The movement of the diaphragm causes the balanced valve V 
at the upper end of the spindle to open or close, as may be necessary 
to maintain the desired lower pressure. Inertia weights T and C pre- 
vent chattering. 

Fig. 526 shows a section through a class G Foster pressure regulator 
or reducing valve. In operation, steam enters at A and passes through 
the main valve port H to the outlet B. Steam at initial pressure passes 
through port C to chamber P and thence to the top of piston T through 
port L, opening the main valve U. Steam at deHvery pressure passes 
through E and raises the diaphragm V against the pressure of spring R, 
allowing spring W to close the auxihary valve X. The pressure in 
chamber J is then equalized by the reduced pressure in ports G and 
the under side of piston X, and thus allows spring Y to close the main 
valve which is then held to its seat by the initial pressure. Any re- 
duction in delivery pressure is transmitted to diaphragm V, and per- 
mits spring to open auxiliary valve X, thereby admitting steam to the 
top of piston T, as previously explained. The delivery pressure is 
adjusted by screw D; thus increasing the tension of spring R increases 
the discharge pressure; and vice versa. The adjustment once made, 
the delivery pressure will remain constant, regardless of any variable 
volume of discharge or of the initial pressure, so long as the latter is 
in excess of the deUvery pressure. TT, Fig. 494, shows the application 
of a reducing valve to an exhaust steam heating system. Live steam 
is led to the valve through pipe A. It will be noted that the pipe 
leading from the valve to the heating system is much larger than the 
high-pressure supply pipe on account of the increase in volume of the 
low-pressure steam. Reducing valves should always be by-passed to 
permit of repairs without shutting down the system. Care should be 
taken in not selecting too large a reducing valve, as the valve Hft is 
very small and the larger the valve the less will be the lift for a given 
weight of flow and consequently the greater the wire drawing and 
erosion of the valve seat. 

365. Foot Valves. — Whenever a long column of water is to be moved 
in either suction or delivery pipe it is customary to place a check valve 
near the lower end of the column to prevent the water from backing 
up when the pump reverses or shuts down. The check valve placed at 
the end of the suction pipe is called a foot valve. Any check valve may 
be used as a foot valve, though practice limits the choice to the disk 
or flap type as illustrated in Fig. 527. To prevent mbbish from de- 
stroying the action, a strainer or screen is generally incorporated with 



776 



STEAM POWER PLANT ENGINEERING 



the body of the valve. A, Fig. 527, illustrates a single-flap, B a multi- 
flxip and C a disk valve composed of a nest of small rubber valves. The 
single-flap are usually made in sizes f to 6 inches, the multi-flap 7 to 
16 inches, and the disk valve in all commercial sizes from f to 36 inches. 






(B) 

Fig. 527. Types of Foot Valves. 



For large sizes, 16 to 36 inches, the multi-disk valve is given preference, 
since a number of the disks may be disabled without destroying its 
operation. 

Blowoff Valves and Sijstems: Prac. Engr., July 1, 1916, p. 565. 

Steam Stop Valves — A Survey of the Field of Design: Sibley Jour., Apr .-May, 1915. 

Nonreturn Stop Valves: Power, Jan. 18-25, 1916, p. 72, 104. 

The Use and Abuse of Globe Valves: Power and Engr., Jan., 1909, p. 10. 

Gate Valves in Steam Pipe Lines: Power and Engr., Feb. 16, 1909, p. 320. 

Types of Check Valves and Their Operation: Power and Engr., July 6, 1909, p. 11. 



PROBLEMS. 

1. Steam at 200 lb. abs. pressure is conducted through a bare pipe 3 in. nominal 
diameter, 500 ft. long. If the temperature of the room is 80 deg. fahr. calculate the 
total heat loss per hour. 

2. If the pipe is covered with a single thickness of "Sall-Mo Air Cell" determine 
the saving in heat. 

3. Determine the conductivity of the covering in Problem 2, per inch of thickness. 

4. Determine the size of steam pipe suitable for a 10,000-kw. steam turbine using 
14 lb. steam per kw-hr., initial pressure 215 lb. abs., back pressure 2 in. mercury, 
superheat 125 deg. fahr., if the pipe is 150 feet long and the pressure drop is not to 
exceed 2.0 lb. per sq. in. per 100 ft. 

5. Saturated steam at 125 lb. abs. initial pressure is flowing at the rate of 20,000 
lb. per hr. through a standard 6-in. pipe, 2000 ft. long. Calculate the probable 
pressure drop. 

6. Determine the initial pressure necessary to deliver 400 gallons of water per 
minute through a 5-in. standard pipe 1500 ft. long, fitted with two right angle elbows 
and one globe valve. The water is to be discharged into an open tank. 

7. How many gallons of water will be discharged through a straight length of 
6-in. standard pipe 10,000 ft. long if the initial pressure is 100 lb. per sq. in., and what 
will be the pressure at the discharge end? 

8. Determine the number and size of safety valves for a 500-hp. boiler designed 
to operate at a maximum load of 300 per cent above rating; boiler pressure 250 lb. 



CHAPTER XVI 

LUBRICANTS AND LUBRICATION 

366. General. — The losses due to the friction of the working part 
of machinery include considerably more than the mere loss of power, 
namely, the depreciation resulting from wear of bearings, guides, and 
other rubbing surfaces, and the expense arising from accidents traceable 
to excessive friction. The power absorbed in overcoming friction varies 
with the type of plant and the character of machinery and is seldom 
less than 5 per cent and often greater than 30 per cent of the total power 
developed. In large central stations these losses approximate 8 per cent 
and in weaving and spinning mills will average as high as 25 per cent. 
(Trans. A.S.M.E., 6-465.) These figures refer to properly lubricated 
plants operating under normal conditions. The proper selection of 
lubricant is therefore a very important problem, since, besides the cost 
of the lubricant itself, the loss in power and in wear and tear to ma- 
chinery is no small item. A change of lubricant may frequently result 
in marked increase in economy of operation. Table 130 gives an idea 
of the saving effected in power by the proper selection of lubricants in a 
number of mills. (Power, May 12, 1908, p. 752.) The net financial 
gain depends, of course, upon the cost of the oil. As a general rule a 10 
per cent reduction in friction horsepower will more than equal the cost 
of lubricants for one year. The lubricants most commonly met with 
in power plant practice are conveniently classified as oils, greases, and 
solids, and are of animal, mineral, or vegetable origin. 

Reference books: Archbutt and Deeley, Lubrication and Lubricants; Redwood 
Lubricants; W. M. Davis, Friction and Lubrication; Gill, Oil Analysis; Robinson, 
Gas and Petroleum Engines; Thurston, Friction and Lost Work; Gill, Engine 
Room Chemistry. 

367. Vegetable Oils. — Except for certain special purposes and for 
compounding with mineral oils these possess lubricating properties of 
little practical value, since they decompose at comparatively low tem- 
peratures and have a tendency to become thick and gummy. The 
vegetable oils sometimes employed are linseed, cottonseed, rape, and 
castor. 

368. Animal Fats. — Many animal fats have greater lubricating 
power than pure mineral oils of corresponding viscosity but are objec- 
tionable on account of their unstable chemical composition. They 

777 



778 



STEAM POWER PLANT ENGINEERING 



decompose easily, especially in the presence of heat, and set free acids 
which attack metals. They are seldom used in the pure state and 
are usually compounded with mineral oils. The animal products used 
in this connection are tallow, neat's-foot oil, lard, sperm, wool grease, 
and fish oil, the first named being the most important. In cyhnder 
lubrication, especially in the presence of moisture, the addition of 2 to 
5 per cent of acidless tallow seems to make the oil adhere better to the 
metal surfaces and increases the lubricating effect, while the proportion 
is so small that ill effects from corrosion or gumming are scarcely per- 
ceptible. Animal and Vegetable Oils, Power, Nov. 3, 1914, p. 636. 



TABLE 130. 



EXAMPLES OF REDUCTION IN FRICTION DUE TO PROPER SELECTION OF 

LUBRICANTS. 


No. of 


Country. 


Plant. 


Mill Oils. 
Test I. 


New Oils. 
Test II. 


Per Cent 
of Trans- 
mission to 
Full Load. 


Power 
Reductions. 


Test. 


Full 
Load, 
LH.P. 


Trans- 
mission, 
LH.P. 


Full 
Load, 
LH.P. 


Trans- 
mission, 
LH.P. 


Test 
I. 


Test 
II. 


Full 
Load, 

Per 
Cent. 


Trans- 
mission, 
Per 
Cent. 


1 
2 A 


America 

America 

America 

America 

England 

England 

England 

England 

Ireland 

Scotland 

Scotland 

Germany 

Germany 

Germany 

Germany 

Russia 

India. . . 


Cotton 

Worsted 

Worsted 

Cotton 

Cotton 

Cotton 

Worsted........ 

Weaving 

Linen 


.543.21 

611.60 

702.90 

786.00 

1408.60 

1428.40 

348.10 

495.00 

110.70 

177.70 

325.10 

263.41 

341.36 

341.36 

1135.20 

1238.80 

642.60 

346.60 

364.70 

465.40 

511.37 

6.74r 

137.80 

84.00 


192.70 

'356:6o' 
357.90 
111.10 
146.60 
49.90 
61.80 
161.40 
114.03 
118.24 
1U.29 
362.60 

'236'. 70' 

"1:772 
74.90 
31.60 


481.75 
596.30 
648.70 
758.00 
1301.80 
1358.70 
327.50 
453.60 
93.10 
164.60 
293.50 
239.35 
290.53 
299.30 
1034.20 
1069.10 
596.80 
313.60 
336.80 
390.40 
482.43 

5.121 
116.00 
65.30 


168.90 


35.4 


35.0 


11.31 
2.50 
7.80 
3 56 
7.60 
4.90 
5.90 
8.40 
15.90 
7.40 
9.70 
9.10 
14.90 
12.30 
8.89 
13.70 
7.10 
9.50 
7.70 
16.20 
5.60 
24.00 
15.80 
22.30 


12.35 


B 










3 










4 A 
B 
5 
6 

7 


319.30 
348.90 
99.50 
127.50 
38.60 
56.10 
147.30 
97.11 
95.67 
119.28 
328.10 


25.3 
25.0 
31.9 
29.6 
45.0 
34.7 
49.6 
43.2 
31.7 
41.3 
31.9 


24.5 
25.7 
30.4 
28.1 
41.4 
34.0 
50.2 
40.5 
32.9 
39.8 
31.7 


10.30 
2.50* 
10.40 
13.00 
22.70 


8A 
B 
9 

10 A 
B 
11 


Woolen 

Woolen 

Cotton 

Worsted 

Worsted 

Jute 


9.20 
8.70 
14.80 
19.10 
15.57{ 
9.51 


12 


Cotton 

Cotton 

Cotton 

Flour 




13 


202.20 


35.9 


33.9 


12.40 


14 


Japan 




15 












16 


England 

Germany 

England 

England 

England 


Paper 










17 


Paper 










18 
19 
20 


Brass shop 

Iron shop 

Wood shop.... 


i.53r 
68.10 
25,40 


26.2 
54.3 
37.6 


29.8 
58.7 
38.8 


13.80 
9.10 
19.60 



* Same oil after nine months' use. 
X = Electrical units. 



t Not full load of miU. 



X Morning load. 



369. Mineral Oils. — These are all products of crude petroleum and 
form by far the greater part of all lubricants. They present a wider 
range of lubricating properties than those derived from animal or 
vegetable sources, the thinnest being more fluid than sperm and the 
thickest more viscous than fats and tallows. They are not easily oxi- 
dized, do not decompose, become rancid, or contain acids. 

Mineral lubrication oils may be classified as 



LUBRICANTS AND LUBRICATION 



779 



(1) .Distilled oils, which are produced b}^ distillation from crude 
petroleum and made pale, amber colored, and transparent by treatment 
with acid and alkali. 

(2) Natural oils, which are prepared from crude petroleum, from 
which grit, suspended and tarry impurities have been removed. They 
are dark and opaque and are rich in lubricating properties. 

(3) Reduced oils, or heavy natural oils, from which the hghter hydro- 
carbons have been evaporated and from which the tarry residue has 
been removed by filtration. 

369a. Solid Lubricants. — Dry graphite, soapstone, and mica are some- 
times used as lubricants, though they are usually mixed with grease or 
oils. They cannot easily be squeezed or scraped from between the sur- 
faces, and are consequently suitable where very great weights have to 
be carried on small areas and when the speed of rubbing is not high. 
The coefficient of friction of such lubricants is high, and when economy 
of power is essential better results may be secured by the use of liber- 
ally proportioned rubbing surfaces and liquid lubricants. Under certain 
conditions of pressure and speed these lubricants will sustain, without 
injury to the surfaces, pressures under which no liquid would work. 



Drops 
per M 

Ko-l Ajto Cylinder Oil (alone) 6 



Press. a 
. Lbs. per oj 

150 411 



|140 
|il20 



H 60 













2 " 


" -f 


1.35 Graphit 


/ 






^ 


L ^^ 




elOU 


« 


/No.l 


No.2, 




& 




/ 4 Kerosene 


" it 


^ 






5 


^ ^ 


_y5^v 


atev 


.. 


/ 


No. 3^^ 














^h==^ 


' No.l' 






lii 








t 1 


No.5^ 






^K ^ 
















n^ 









8 70 148 



30 60 30 60 30 60 30 60 

Time in Minutes 

Fig. 528. Tests of Graphite Mixed with Various Lubricants. 



Deflocculated graphite suspended in oil or water, and designated com- 
mercially as ''oildag" and '^aquadag" respectively, is finding favor 
with many engineers. Graphite in this deflocculated condition remains 
suspended indefinitely in water and oil, readily adheres to the journal, 
has great wearing properties, and is easily appfied to the wearing sur- 
faces. From numerous and long-continued trials it appears that 0.35 
per cent serves adequately for all purposes. Temperature curves of 
deflocculated graphite in combination with various carrying fluids are 
given in Fig. 528. For further data pertaining to the curves in Fig. 528 
and for an extensive discussion on the subject of lubrication consult Lu- 
hrication and Lubricating, by C. F. Maberg, Jour. A.S.M.E., Feb. and 
May, 1910. 



780 



STEAM POWER PLANT ENGINEERING 



370. Greases. — Under this name may be included the various com- 
pounds which consist of oils and fats thickened with sufficient soap to 
form, at ordinary temperatures, a more or less solid grease. Those 
usually employed are lime, soda, or lead goaps, made with various fats 
and oils. ''Engine" greases are thickened with a soap made from 
tallow or lard oil and caustic soda, and often contain neat's-foot oil, 
beeswax, and the hke. For exceptionally heavy pressures, graphite, 
soapstone, and mica are sometimes added to the grease. Table 131 
gives an idea of the characteristics of a number of greases. (Prac. 
Engineer, U. S., Apr. 1911, p. 293.) The friction tests were made on 
a small Thurston oil testing machine, 320 r.p.m. and bearing pressure 
of 240 pounds per square inch of projected area. These results are 
purely comparative under the given conditions of rubbing surfaces, 
speed and pressure. For results of these greases tested on a large 
Olsen oil machine consult reference given above. 

Commercial Lubricating Greases: Prac, Engineer, U.S., Apr., 1911, p. 293; Tests 
of Grease Lubrication, Ibid., p. 295; Am. Mach., Aug. 24, 1911, p. 356; Power, Nov. 
8, 1910, p. 1998. 



TABLE 131. 

LUBRICATING CHARACTERISTICS OF A NUMBER OF GREASES. 



Type. 


Class. 


Melting 
Point. 
Deg. F. 


Per Cent 
Soap. 


Kind of 
Soap. 


Per Cent 
Free Acid 
as Oleic. 


Average 
Coefficient 
Friction. 


A Mineral 

B Mineral 

C Mineral 


Summer 

Summer 

Winter 

Winter 

Winter 

Winter 

Summer 


167 

178 
165 
163 
142 
125 
120 
41 


38 

20 

23 

16 

19 
1.4 
2.1 



Lime 

Lime 

Lime 

Lime 

Lime 

Potash 

Potash 


Trace 

0.3 

6 1 


Trace 






^ 075 
054 
063 


D Mineral 

E Mineral 

F Tallow No. 3 . . . 
G Tallow No. XX 
H Lard oil 


0.057 
0.046 
0.022 
0.029 
0.011 













Type. 


Final Coefficient 

Friction After 

3-Hr. Run. 


Maximum Temper- 
ature of Bearing 
Above that of 
Room, Degs. F. 


Final Temperature 
of Bearing Above 
that of Room at 
End of 3-Hr. Run. 
Degs. F. 


A Mineral 


0.075 

0.050 
0.063 
0.054 
0.046 
0.012 
0.018 
0.010 


70 

70 
76 
69 
58 
38 
45 
13 


68 


B Mineral 


58 


C Mineral .... 


65 


D Mineral 


58 


E Mineral 


50 


F Tallow No. 3 


18 


G Tallow No. XX 


32 


H Lard oil. 


12 







I 



LUBRICANTS AND LUBRICATION 781 

371. Qualifications of Good Lubricants. — A good lubricant should 
possess the following quahties: 

(1) Sufficient ''body" to prevent the surfaces from coming into con- 
tact under conditions of maximum pressure. 

(2) Capacity for absorbing and carrying away heat. 

(3) Low coefficient of friction. 

(4) Maximum fluidity consistent with the "body" required. 

(5) Freedom from any tendency to oxidize or gum. 

(6) A high ''flash point" or temperature of vaporization and a low 
congealing or "freezing point. ^^ 

(7) Freedom from corrosive acids of either metalhc or animal origin. 

372. Testing Lubricating Oils. — There is no question but that the 
lubricant best suited for a given set of conditions can only be determined 
by an actual practical test under service conditions. Each plant is 
an individual problem since certain grades and qualities of oil which 
work perfectly in some cases have proved entirely unsatisfactory in 
others where the conditions appeared to be exactly the same. Neverthe- 
less, in order to avoid needless experiment and to limit the number of 
acceptable lubricants to a minimum it is desirable to know certain 
characteristics which will indicate whether or not the particular lu- 
bricant under consideration is unfitted for the desired service. The small 
consumer must depend upon the reputation of the concern from which 
he is buying for reliable data pertaining to the qualifications of their 
products, since the cost of conducting a series of preliminary or identi- 
fication tests is out of all proportion to the actual cost of the lubricant. 
The large consumer on the other hand may find it to be worth while 
to conduct an elaborate series of tests before drawing up contracts for 
the oil supply. 

The complete test of an oil consists of three parts: Chemical, physi- 
cal, and practical. 

373. Chemical Tests of Lubricating Oils. — To pass the chemical 
tests of the Navy Department * "all oils should be neutral in reaction 
and should not show the presence of moisture, matter insoluble in 
petroleum ether (hard asphalt), matter insoluble in ether alcohol 
(soft asphalt), free sulphur, charring or wax-like constituents, naphthenic 
acids, sulphonated oils, soap, resin or tarry constituents, the presence 
of which indicates adulteration or lack of proper refining. Except in 
oil for engines without forced lubrication, no traces of fixed oils (animal 
or vegetable fats) should be found. 

* Lubricating Oils. Lieut. J. L. Kauffman, U.S.N. , Jour. Am. Soc. Naval Engrs., 
Aug., 1916, p. 692. 



782 STEAM POWER PLANT ENGINEERING 

"In lubricating oil for main engines without forced lubrication, 
approved fixed oils, such as rapeseed, olive, tallow, lard and neat's-foot 
oil, may be used. When the foregoing fixed oils are used, they must 
be well refined with alkalies, unadulterated, containing a minimum of 
free fatty acids, with no moisture or gumming constituents. Ohve 
oil should not have a high specific gravity. If satisfactory emulsifying 
results can be obtained with straight mineral oils on engines without 
forced lubrication, they may be submitted for service test." 

The most satisfactory procedure is to have the various tests made 
by a competent chemist but since a number of plants are provided with 
the necessary equipment the tests stipulated by the Navy Department, 
and which are representative of current commercial practice, will be 
described in a general way. 

Moisture. — Heat 3 to 4 cc. in a test tube (the walls of which have 
been thoroughly wet with oil) in a bath of liquid paraffin up to 300 
deg. fahr. Oils containing water will form emulsions on the walls and 
cause foaming and spluttering. A test is also made with a mixture 
of oil and eosin to determine faint traces of moisture by changes of color. 
The presence of moisture is particularly undesirable in transformer 
oils, but there is danger of its forming objectionable emulsions in any 
straight mineral oil. 

Sulphur. — Boil about 50 cc. of oil with a piece of bright metallic 
sodium for half an hour; add water, heat and stir until the sodium is 
dissolved; pour off the water and test the remainder with a fresh 1 
per cent solution of sodium nitroprusside. If the mixture turns 
violet color, the oil contains sulphur. When sulphur is found, an 
additional test for sulphonated oils is made. 

Acids or Alkalies. — Heat for one-half hour with frequent stirring 
25 cc. of oil and 50 cc. of neutral distilled water. Test a few cubic 
centimeters of the mixture first with methyl-orange to determine the 
acids, and another portion with phenolphthalein for the determination 
of alkahes. Acids and alkalies cause emulsions. Acids also cause 
corrosion of journals and other metal parts. 

Matter Insoluble in Ether Alcohol. — Shake 11 cc. of oil and 14 cc. of 
ether alcohol (8 parts ether and 6 parts alcohol). After standing 12 
hours, note the precipitate, if any, at the bottom of cylinder. The 
precipitate will be asphalt, and even a trace would make the oil un- 
desirable as a lubricant. Asphalt would cause scoring of journals and 
clogging of oil lines. 

Matter Insoluble in High-grade Gasoline. — Shake 2 cc. of oil and 
about 300 cc. of high-grade gasoline (86-88 Baume gravity). After 
standing 12 hours, note precipitate, if any, in the bottom of glass. 



LUBRICANTS AND LUBRICATION 783 

The precipitate will be soft asphalt or carbon particles, and a shght 
trace would make the oil undesirable. 

Tarnj or Suspended Matter. — Same as the foregoing, except using 
5 cc. of oil and 95 cc. of gasoline and allowing it to stand for half an hour; 
then examine deposit, if any, for dirt or tarry matter. 

To Detect Fixed Oils. — Heat 10 cc. of oil with a small piece of metal- 
lic sodium. If the mixture becomes gelatinized or a semisolid, it in- 
dicates the presence of fixed oils. If an equal volume of oil is heated 
alone to the same temperature, the viscosity of the two samples can be 
compared; if the oil contains fixed oils (animal or vegetable oils), the 
sample with sodium will be much heavier than the sample heated alone. 

Effect of Heat. — Heat 5 cc. of oil in test tube over flame until vapors 
are evolved and compare the color of the heated oil with that of unheated 
oils. If the heated oil turns black, it shows the presence of undesirable 
carbon or hydrocarbons. 

Gumming Test."^ — This is particularly appHcable to petroleum oils 
and is used to indicate the extent to which the oil has been refined. It 
serves indirectly to indicate the extent to which the oil may be expected 
to change due to oxidation when in use. Numerous opportunities have 
been offered to check the results obtained with this test and results ob- 
tained in practice with the same oils, and all of this experience tends to 
show the great value of the gumming test. 

This test is made by putting a small quantity of the oil to be tested 
in a small glass vessel, such as a cordial glass, and then mixing with it 
an equal quantity of nitrosulphuric acid. A properly refined oil will 
show httle, if any, change, but a poorly refined oil will be indicated by 
the separation of large quantities of material of dark color. This color 
is due to the oxidation of the tarry matter contained in the lubricant. 
Experience has shown that oils containing large percentages of tar ab- 
sorb the most oxygen, that is, they are mildly drying oils. 

The results obtained by the gumming test agree well with carbon- 
residue tests made by distilling to dryness in a glass or a fused quartz 
flask. The carbon-residue test has been found of great assistance in 
choosing a satisfactoiy cylinder lubricant for gas engines, as a large 
amount of carbon means trouble in the engine cylinder. The lowest 
carbon content mentioned by the author was 0.11 per cent. The oil 
giving this test showed no tarry matter when tested with nitrosulphuric 
acid. In general, a gas-engine oil should not contain more than 0.5 
per cent carbon as determined by the carbon-residue test. 

374. Physical Tests of Lubricating OUs. — The physical characteris- 
tics usually involve (1) color; (2) odor; (3) specific gravity; (4) flash 

* Prof. A. H. GiU. 



784 STEAM POWER PLANT ENGINEERING 

point; (5) fire point; (6) cold point; (7) viscosity; (8) emulsion; 
(9) evaporation; and (10) friction. The following tests, unless other- 
wise indicated, refer specifically to the requirements of the Navy De- 
partment which, as previously stated, are representative of current 
commercial practice. 

Color. — The color, although having no influence on the lubricating 
value, may be used to identify the sample. American oils fluoresce with 
a grass-green color. Russian oils have a blue sheen; oils containing 
distillation residues and unfiltered oils are brown to green-black in 
reflected light. Nearly all mineral machinery oils are distilled and 
filtered to some extent and are transparent in a test tube, the colors 
ranging from a yellowish white to a blood red. The color may be 
determined in a tinctometer by comparing with different-colored 
glasses or lenses. These glasses are numbered and for machinery oil 
extend from No. 1 (white) to No. 6 (red). 

Odor. — The odor may be determined by heating in a test tube or 
by rubbing on the hand, by which means fatty oils, coal tar, rosin oils, 
etc., may be detected. 

Specific Gravity. — The specific gravity is obtained by the use of the 
''pyknometer, " this term signifying any vessel in which an accurately 
measured volume of liquid can be weighed. The bottle is first filled 
with distilled water at a temperature of 60 deg. fahr., and the weight 
of the water determined. The bottle is then filled with oil at a tempera- 
ture of 60 deg. fahr. and the weight of the oil determined. The weight 
of the oil divided by the weight of the water gives the specific gravity at 
60 deg. fahr. The Baume gravity is obtained by using the Baume 
hydrometer, which is simply an ordinary hydrometer with a certain arbi- 
trary scale. Baume gravity may be converted into specific gravity by 
the following formula: 140 

Sp.gr. = i30 4_Bau"^* 

Baume gravity is largely used in commercial practice. 

The specific gravity does not affect the lubricating value of an oil, 
but indicates to the experienced oil man the locality from which the crude 
oil is obtained. For instance, the specific gravities of the lubricating 
oils tested at the Experiment Station vary from 0.864 to 0.945. A 
Baume gravity of 32 corresponds to a specific gravity of 0.864, and a 
Baume gravity of 18.1 to a specific gravity of 0.945, so that an increase 
in specific gravity is a decrease in Baume gravity. The paraffin-base 
oils of Pennsylvania derivation have an average specific gravity of 
0.875 with a corresponding Baume gravity of 30. The asphaltic-base 
oils from Texas and CaUfornia have an average specific gravity of 
0.930 with a corresponding Baume gravity of 20. 



LUBRICANTS AND LUBRICATION 



785 



TABLE 132. 
SPECIFIC GRAVITY AND GRAVITY BAUME OF A NUMBER OF LUBRICANTS. 



Water 

Cylinder oil 

Cylinder oil 

Heavy engine oil. . . 
Medium engine oil. 
Light engine oil. . . . 
Castor machine oil. 

Lard oil 

Sperm oil 

Tallow oil 

Cottonseed oil 

Linseed oil 

Castor oil (pure) . . . 

Palm oil 

Rape-seed oil 

Spindle oil 



Specific Grav- 


Gravity 


Flash Test. 


ity. 


Baumd. 


Degrees F. 


1.000 


10 
24.5 




.9090 


575 


.8974 


26 


540 


.9032 


25.5 


411 


.9090 


24 


382 


.8917 


27 


342 


.8919 


27 


324 


.9175 


23 


505 


.8815 


29 


478 


.9080 


24.5 


540 


.9210 


22 


518 


;9299 


19 


505 


.9639 


15 
25 




.9046 


405 


.9155 


23 
33 




.8588 


312 



Flash Point. — The flash point is determined with both the Cleveland 
open cup and the Pensky-Martin closed cup. The flash point of all 
oils is determined as a measure of their volatility. The flash point of 
steam-cylinder oils is of primary importance, the required flash point 
depending on the temperature of the steam at the engine. With lu- 
bricating oils for bearings the flash point is important only in that it 
indicates the volatility of the oils and the presence of kerosene or naph- 
tha fractions, with the accompanying fire risks. In the case of very low 
flash-point lubricating oils, it is desirable to run a special distillation 
or volatility test, mentioned under chemical tests. The flash point 
determined with the open cup is higher than with the closed cup, as 
the inflammable gases on the surface of the oil are disturbed by the air 
currents in the open cup. These differences range from 5 deg. to 40 deg. 
with the average at 20 deg. The presence of very light ends (kerosene, 
naphtha, etc.) may increase this difference to 100 deg. 

Fire Point. — This is the temperature at which the oil burns and is 
determined by raising the temperature about 3 deg. a minute, applying 
the flame for about a second. The fire, or burning, point is from 30 deg. 
to 65 deg. higher than the flash point with all lubricating oils, the light 
oils having a difference of about 40 deg. 

Cold Point. — Mineral oils become more viscous on cooling, and 
finally solidify. In lubricating oils refined from paraffin-base crudes, 
cooling first causes the parafl&n particles to soUdif}^ which gives the 
oil a cloudy appearance; with this class of oils this change is known as 
the cloud point. 



786 STEAM POWER PLANT ENGINEERING 

The Committee on Lubricants of the American Society for Testing 
Materials uses the words ''cold test" as a general term, with subheads 
of ''cloud test" and "pour test." The method recommended by this 
committee is used at the Experiment Station, and in substance is as fol- 
lows: Heat the oil to 150 deg. fahr. and cool by air to 75 deg. fahr. 
Take a bottle about IJ in. inside diameter and 4 to 5 in. high and pour 
in oil to a height of IJ in. from the bottom. Insert a cold-test ther- 
mometer (specially made, using colored alcohol, and with a long bulb) 
through a tight-fitting cork. A special jacket is used having an inside 
diameter about J inch larger than the bottle. Ice or any other cooling 
medium is packed around this jacket. When the oil is near the expected 
cloud point, at every 2 deg. drop in temperature remove the bottle and 
inspect the oil, being careful not to disturb the oil. When the lower 
half becomes opaque, read the thermometer; this reading is taken as 
the cloud point. The cold, or pour, test is simply a continuation of 
the cloud test, except that the temperature is noted every 5 deg. and 
the bottle tilted till the oil flows. When the oil becomes solid and will 
not flow, the previous 5-deg. point is taken as the cold point of the oil. 

Viscosity. — The viscosity of a lubricating oil is the most important 
factor to be determined. The viscosity of an oil is inversely propor- 
tional to its fluidity and is a measure of its internal friction or resistance 
to flow. Viscosity is sometimes called "body" and is determined by 
a viscosimeter. There are a number of different instruments for this 
purpose but no recognized standard instrument or method, so that 
"viscosity" conveys no meaning unless the name of the instrument, 
the temperature, and the amount of oil tested are given. Nearly all 
instruments are of the orifice type; that is, the ^dscosity of an oil is 
taken as the number of seconds required for a given amount to flow 
through an orifice at a given temperature. By "specific viscosity" is 
meant the ratio of the time required for the oil to run out to that of an 
equal quantity of water at 60 deg. fahr. The viscosity of engine oils 
is usually taken at 100 to 130 deg. fahr. and of cylinder oils at 210 deg. 
fahr. The absolute viscosity is determined from the amount flowing 
through capillary tubes, the results being given in C. G. S. units. The 
determination of the absolute viscosity is a veiy difficult operation 
requiring complex apparatus and a relatively long time. Several ab- 
solute viscosimeters have been invented; but to date none of them is 
considered practical enough for the routine testing of oil. 

The accepted theory advanced by Ubhelohde * is that the absolute 
viscosity is directly proportional to the internal friction of the lu- 
bricant, and that the viscosity is a direct indication of the friction 

* General Electric Review, November, 1915. 



I 



LUBRICANTS AND LUBRICATION 787 

developed in a l)caring;. If Ubhelohde's conclusion is substantiated 
a very great advance will have been made and it will be possil)le to 
duplicate any friction results by duplicating the viscosity of the lubri- 
cant. 

In general^ the lower the viscosity the lower will be the friction, 
but since the rubbing surfaces should have as much lubricant between 
them as possible it is necessary to have sufficient viscosity to prevent 
them from '' seizing. '* Under normal conditions of bearing lubrication 
the lightest oil that will prevent seizing should be used to obtain a 
minimum frictional loss. 

Viscosity and its Relation to Lubricating Values: Power, Jan. 11, 1916, p. 37. 

Emulsion Tests. — Emulsion tests are made on all straight mineral 
oils except cylinder oils. Four emulsion runs are made, using 40 cc. of 
oil in each case and (a) 40 cc. of distilled water; (6) 40 cc. of salt water; 
(c) 40 cc. of normal caustic-soda solution; (d) 40 cc. of boiling distilled 
water. The mixture is stirred with a paddle for five minutes at 1500 
revolutions per minute and is kept at a temperature of 130 deg. fahr. 
during the stirring and while separating. On oils used with forced 
lubrication or on ice machines, the oil must completely separate from 
the mixture in less than 20 minutes. The emulsion is made with dis- 
tilled and salt water, and a normal caustic-soda solution is also taken, 
as there is a possibility of water containing boiler compound getting into 
the system. Boiling distilled water is used in case gland steam or 
water runs into the oil system. These emulsion tests are considered of 
the greatest importance, as an oil on any type of forced lubrication sys- 
tem must not emulsify. If emulsions do occur, it will mean clogging of 
the oil lines, forming of residues in the base of the bearings, with a re- 
sultant loss of a large amount of oil. 

Evaporation Tests. — It is advisable to include an evaporation test 
with the flash test of lubricants. The evaporation test is made by 
exposing about 0.2 gram of oil at a proper temperature and determining 
the loss by weight in a given time. 

Friction Tests. — The coefficient of friction as determined from fric- 
tion-testing machines is useful in obtaining a comparison of oils under 
the test conditions, but gives little information concerning the action of 
the oil under the widely different conditions found in actual practice. 

Table 133 gives the physical properties of a number of lubricating 
oils, with their particular fields of application. 

375. Service Tests. — These tests are the real proof of the commercial 
value of the lubricant for a given service. The lubricant is tested 
under actual operating conditions and that one selected which gives 



788 



STEAM POWER PLANT ENGINEERING 



TABLE 133. 
PHYSICAL CHARACTERISTICS OF A NUMBER OF LUBRICANTS. 

(Power, December, 1905, p. 750.) 



Kind of Oil. 


Use and Adaptation. 


o 


^"1 

8^ 






Viscosity 
at 70 De- 
grees. 


High-pressure cylinder 
oil. 


For steam cylinders using dry 
steam at pressures from 110 
to 210 pounds. 


25 

to 

24.5 


30 


600 
to 
610 


645 
to 
660 


175 

to 

205 


General cylinder oil . . 


For steam cylinders using dry 
steam at 75 to 100 pounds. 
For air compressor cylinders 
when made from steam-re- 
fined mineral stock and when 
viscosity is 200. 


26 

to 

25.5 


30 


550 
to 

585 


600 
to 
630 


180 
to 
190 


Wet cylinder oil. 
(Remark 1.) 


For use where the steam is moist, 
especially in compound and 
triple expansion engines. 


25.8 

to 
25.3 


30 


560 
to 
585 


600 
to 
630 


150 
to 

185 


Gas engine cylinder oil. 
(Remark 2.) 


For gas engine cylinders. Neu- 
tral mineral oil compounded 
with an insoluble soap to give 
body. 


26.5 


30 


320 


350 


300 


Automobile gas engine 
oil. (Remark 3.) 


For automobile gas engines and 
similar work. 


29.5 


30 


430 


485 


195 


Heavy engine and 
machinery oils. 


For heavy slides and bearings, 
shafting, and horizontal sur- 
faces. 


30.5 

to 
29.5 


30 


400 


440 
to 
450 


170 
to 
195 


General engine and 
machine oils. 


For high-speed dynamos and 
machines. 


30.8 
to 
30 


30 


400 
to 
420 


450 
to 
470 


175 
to 
190 


Fine and light machine 
oils. 


For fine work, from printing 
presses to sewing machines 
and typev^Titer oils. With a 
cold test of 25° to 28° and a 
viscosity of 140° this makes 
an excellent spindle oil. 


32.5 

to 
30.2 


30 


400 


440 


110 
to 
160 


Cutting and heat dis- 
sipating oils. 
(Remark 4.) 


For cutting tools, screw cutting 
and similar work. 


27 
to 
23 


30 


410 
to 
420 


475 
to 

480 


210 
to 

175 


Refrigerating oils 


For ice machinery 


30.2 





200 


225 


165 








Wet service and marine 
oils. (Remark 4.) 


For marine service, or where a 
great deal of moisture must 
be handled. 


28 


1 

30 


430 


475 


230 


Greases , 


They are used in special work 
and for heavy pressures mov- 
ing at slow velocities. 

















Remark 1. — May contain not over 2 to 6 per cent of refined acidless tallow oil in the high- 
pressure oils and not over 6 to 12 per cent in the low-pressure oils. 

Remark 2. — The reason for using an insoluble soap such as oleate of aluminum is that it 
is impossible to decompose the soap with a high heat ; the soap, although not a lubricant, is a 
vehicle for carrying some oil. 

Remark 3. — Owing to a lack of body, this oil will not interfere with the sparking by depos- 
iting carbon on the platinum point. 

Remark 4. — May contain 30 to 40 per cent of pure strained lard oil. 



I 



LUBRICANTS AND LUBRICATION 789 

the best overall economy, such factors as first cost,, quantity used, effect 
on the rubbmg surfaces, maintenance and attendance being taken into 
consideration. Ha\'ing determined the particular grade of lubricant 
which gives the best returns the tests previously mentioned are made 
and the results incorporated in the specifications so as to insure delivery 
of that particular grade of lubricant. Large consumers frequently 
employ the services of an experienced lul^ricating engineer under the 
supervision of the plant engineer or millwright for determining the 
lubricant best suited for the different classes of machinery. 

Testing of Lubricating Oils: Power, Apr. 13, 1915, p. 522, 

376. Atmospheric Surface Lubrication. — In a general sense all jour- 
nals, slides, and ''atmospheric" surfaces should be lubricated with 
straight mineral oils (as free from paraffin as possible), except when in 
contact with considerable water, in which case it is advisable to add 
20 to 30 per cent of lard oil. Vegetable, oils, paraffin oils, and animal 
oils (except lard oil as above stated) are not recommended for general 
engine and dynamo service. The test requirements of a number of 
classes of lubricants are outfined in Table 133 and represent current 
practice. Bearings, guides, and all external rubbing surfaces may be 
lubricated in a number of ways. (1) They may be given an inter- 
mittent appfication of oil, as, for example, with an oil can; (2) they may 
be equipped with oil cups with restricted rates of feed; and (3) they 
may be flooded with oil. The relative lubricating values of the systems 
have been estimated approximately as follows (Power, December, 1905, 
p. 750): 



Intermittent. . . . 
Restricted feed. 
Flooded bearing, 



CoeflBcient of Fric- 
tion. 



0.01 and greater 
0.01 to 0.012 
0.00109 



Comparative 
Value. 



72 and less 
79 to 86 
100 



377. Intermittent Feed. — Intermittent applications are ordinarily 
Hmited to small journals, pins, and guides which are subject to light 
pressures and which do not easily permit of oil or grease cups, as, for 
example, parts of the valve gear of a Corliss engine, governors, and link 
work. On account of the labor attached and the frequent doubt about 
the oil reaching the wearing surfaces this method of lubrication is 
hmited as much as possible even in the smallest plants. 

378. Restricted Feed. — In the average power plant the major part 
of the lubrication is effected by means of oil cups which are filled at 



790 



STEAM POWER PLANT ENGINEERING 



intervals by hand or by mechanical means, the oil being fed from the 
cup by drops, according to the requirements. 

379. Oil Bath. — In large power plants the principal journals and 

wearing parts are supplied -with a 
continuous flow of oil which com- 
pletely ''floods" the rubbing sur- 
faces. The oil is forced to the vari- 
ous parts either by gravity from an 
elevated tank or by pressure from 
a pump. After the oil leaves the 
bearings it flows into collecting 
pans, thence into a receiving and 
filtering tank, and finally is pumped 
back into an elevated reservoir and 
used over and over again. The 
little lost by leakage and deprecia- 
tion is replenished by the addition 
of new oil to the system. 

380. OU Cups. — Fig. 529 illus- 
trates the application of sight-feed oil cups to the crosshead and slides 
of a reciprocating engine. The oil is fed into the cups by hand and 
gravitates to the rubbing surfaces, the rate of flow being regulated by 




Fig. 529. Oil-cup Lubrication, 
Hand-fiUed. 





Fig. 530. Nugent 's Telescopic Oiler. 

a needle valve. Cups A and B feed directly to the crosshead guides, 
but the oil from cup D flows to the bottom orifice 0, from which it is 
wiped by a metalhc wick S, and carried by gravity to the wrist pin. 



LUBRICANTS AND LUBRICATION 



791 



zzzzzzzMzzzn 




Fig. 531. Oil-ring Lubrication. 



381. Telescope Oiler. — Fig. 530 shows the application of a telescopic 
oiler to a crosshead and guides. and C are sight-feed oil cups, the 
former feeding directly to the top guide through the tube S. The oil 
from C flows by gravity through the swing joint into the telescopic tubes 
P, R, and thence to the pin through the lower swing joint as indicated. 
As the crosshead moves back and 
forth, the pipe P shdes into and 
out of pipe R, the oil being thus 
conducted directly to the pin with- 
out wasting. A device of this type 
installed on a high-speed automatic 
engine at the Armour Institute of 
Technology has been in operation 
for five years without cost for re- 
pair or renewal. 

382. Ring Oiler. — Small high- 
speed engines are often oiled by 
the oil-ring system, as illustrated in 
Fig. 531. The shaft is encircled by 
several loose rings which dip into a 
bath of oil in the base of the pedestal or frame and, rolling on the shaft 
as it turns, carry oil to the top of the shaft where it spreads to the 
bearings. In some cases the rings are replaced by loops of chain. 

Ring Lubrication: Power, Jan. 9, 1917, p. 42. 

383. Centrifugal Oiler. — Fig. 532 illustrates the application of a 
centrifugal oiler to a side-crank engine. The oil supply is regulated by 

the sight-feed cup C and flows by gravity 
to the pipe P in line with the center of 
the crank shaft. Centrifugal force throws 
the oil outward through pipe B to the 
center of the pin D, which is drilled longi- 
tudinally and radially so as to distribute 
the oil upon the bearing surface. 

384. Pendulum Oiler. — Fig. 533 illus- 
trates the application of a pendulum oiler 
to the crank pin of a center-crank engine. 
Oil cups and pendulum P are fastened to 
the crank shaft S by trunnion T. The pendulum holds the cup ver- 
tical, since the friction of the trunnion is not sufficient to revolve it. 
Oil flows along the center of the crank shaft under the head of oil in 
cup and is thrown outward to bearing B by centrifugal force. 




Fig. 532. Centrifugal Oiler. 



792 



STEAM POWER PLANT ENGINEERING 



385. Splash Oiling. — In some high-speed engines tlie crank, con- 
necting rod, and crossheads are inclosed by a casing, the bottom of 
which is filled with oil to such a depth that at each revolution of the 




Fig. 533. Pendulum Oiler. 



crank, the end of the connecting rod is partly submerged. The result 
is that the oil is splashed into every part of the chamber, and the crank 
pin, crosshead pin, and crosshead slides practically run in an oil bath. 



Gauge 



Over Flow 




By-pass 



To Waste 

Basement Floor Line 



Fig. 534. Simple Gravity Feed System. 

386. Gravity Oil Feed. — Fig. 534 illustrates a simple gravity oil-feed 
system. The oil to the engine is supplied from the oil tank by pipe D 
under pressure corresponding to the height of the tank above the oil A 
cups. After performing its function the oil gravitates to the filter and " 



LUBRICANTS AND LUBRICATION 



793 



from the latter to the oil reservoir, from which it is pumped back to 
the supply tank, the overflow being returned to the reservoir through 
pipe N. Operation is interrupted only when new oil is to be added to 
the system from the barrel through the flexible filling pipe. In case 
the oil tank is put out of commission, or the supply pipe becomes clogged, 
full pump pressure may be used by closing valves R and S and opening 
valve E. The make-up oil is small in amount compared to. the quantity 
circulated. The reclaiming and purifying of the oil are essential if the 
bearings are to be flooded, otherwise the cost of oil would be prohibitive. 
At the power house of the South Side Elevated Railway the daily 
circulation (24 hours) of engine oil is approximately 1500 gallons. The 
make-up oil amounts to eight gallons. 

An objection sometimes made to the above system is that the varying 
heights of oil in the supply tank may cause considerable variation in 
pressure at the oil cups, causing them to feed faster when the tank is 
full and slower when the tank is nearly empty. This applies only to 
installations where the supply tank is filled intermittently, 



Inlet 



,^A.\r Hole 



z: 



Small Brass Pipe 



"ti^t-^sr 



I 



Fig. 535. Low-pressure Gravity Feed, Constant Head. 



387. Low-pressure Gravity Feed. — Fig. 535 shows the application of 
a low-pressure oiling system in which the level in the sight feeds is 
kept constant. A is the main supply tank, B^ and B^ the upper and 
lower gauges indicating the oil level, C the supply pipe running to the 
engines, and D a small standpipe closed at one end and vented near the 
top. The reservoir is supplied with oil by the valve marked '4nlet. " 
When the tank is fiilled the oil rises in the standpipe D a corresponding 



794 



STEAM POWER PLANT ENGINEERING 



height. The inlet valve is then closed and the oil in the standpipe feeds 
down to the level of the sight feeds or to a point where the air will enter 
the bottom of the tank. This will be the constant oil level, since oil 
flows from the tank only in proportion to the amount of air admitted. 
A head of 6 inches has been found to give the best results. (Engineer, 
U. S., March 16, 1903, p. 243.) 

388. Compressed-air Feed. — Fig. 536 shows diagrammatically the 
arrangement of the oiUng system at the First National Bank Building, 
Chicago. The storage tank containing the supply of engine oil is 
under air pressure^ at all times except during the short periods when it 
is being filled with oil from the filter. The air pressure on the surface 




Engine 



Fig. 536. Oiling System at the Power Plant of the First National Bank 
Building, Chicago. 

of the oil forces it to a manifold on the engine from which it is dis- 
tributed to the various oil cups. The oil flows from the different 
bearings to the returns tank located at the base of the engines. When 
the tank is filled air pressure is admitted and the oil forced to the settHng 
tank, which has a capacity of about 400 gallons and is located near the 
ceihng. The oil is allowed to settle and the entrained water and foreign 
material are drained to waste. The oil gravitates from this tank to a 
series of Turner oil filters. When a new supply of oil is needed, valves 
A and B are closed and vent valve C opened, cutting off the supply of 
air and reducing the pressure to atmospheric. Valve D is then opened 
and oil flows from the filters to the storage tank. 

389. Cylinder Lubrication. — The test requirements for cylinder oils 
are outlined in Table 133, from which it will be seen that pure mineral 
oil fulfils practically all requirements for dry steam. In connection 
with moist steam, as in the low-pressure cylinders of compound engines, 
an addition of from 2 to 5 per cent of acidless tallow oil is recommended. 



LUBRICANTS AND LUBRICATION 



795 



Vegetable oils, beeswax, lard oil, degras (wool grease), and the like 
should never be used in compounding cylinder oils. The best cylinder 
oils are made from Pennsylvania stock. For data pertaining to the 
amount and grade of cylinder oil used in a large number of piston engine 
plants see Table I, p. 824, Jour. A.S.M.E., May, 1910. See also ^^Lubri- 
cants and Lubrication," by Dr. C. F. Mabery, Jour. A.S.M.E., Feb., 
1910. 

Cylinder oils must be forced to the parts requiring lubrication against 
the prevailing steam pressure, which is ordinarily accomplished by 
(1) cylinder cups, (2) hydrostatic lubricators, or (3) hand- or power-driven 
force pumps. 

390. Cylinder Cups. — A cylinder oil cup consists essentially of a 
steam-tight brass vessel fitted at the bottom with a pipe connection 
and valve. A screwed cap offers a means of introducing the lubricant 
into the cup. After the cap is in place the valve is opened and the 
cup is subjected to full steam pressure. The pressure in the cup, being 
equal to that in the steam chest or cylinder, permits the lubricant to 
gravitate through the valve into the cylinder. 

Fig. 537 shows a section through an improved 
form of oil cup in which the oil feeds from the 
top instead of the bottom as is the case with 
the common form of cylinder cup. The vessel 
is attached to the steam chest or to the supply 
pipe below the throttle valve. Steam is ad- 
mitted through opening B and, condensing, 
settles through the oil to the bottom. This 
raises the level of the oil until it begins to 
overflow down the same passage by which the 
steam enters. This action is intensified by the 
fluctuation in steam pressure. The rate of 
feeding is regulated by valve C and tested by 
unscrewing plug F. If oil appears through 

opening G, the cup is feeding oil; if steam or water is emitted the 
cup is empty. The cup is filled by means of plug E and the water 
drained at D. 

391. Hydrostatic Lubricators. — The most common method of cylin- 
der lubrication is by means of hydrostatic lubricators of the sight-feed 
class. Fig. 538. The principle of operation is as follows: The lubri- 
cator is filled with cylinder oil by removing cap K, the height of oil 
appearing in glass L. If water is present the oil floats on top as indi- 
cated. After the cap is screwed in place the valves in the condenser 
pipe are opened, subjecting the oil in the vessel to steam-pipe pressure. 




Fig. 537. Leyland Auto- 
matic Cylinder Cup. 



796 



STEAM POWER PLANT ENGINEERING 



Steam is condensed in pipe C, filling tube B and part of C, thus adding 
to the steam pressure the pressure due to the weight of the water column. 
Valve F, which communicates with the top of the vessel by means of 
tube A, is opened wide, as is also the regulating valve 7. The pressure 
at B being greater than that at A by an amount equivalent to the 
height of the water column, forces the oil through A and the '^ sight 
feed" S to the steam pipe. The rate of flow is controlled by the 




STEAM 
PIPE 



Fig. 538. Common Hydrostatic 
Lubricator. 




Fig. 539. Lunkenheimer Sight-feed 
Lubricator. 



regulating valve I. As the oil flows from the vessel its space is occu- 
pied by condensed steam, the height of oil and water being visible in 
glass L. Owing to the small capacity of the lubricator it must be 
refilled frequently. To reduce the amount of labor required with the 
above apparatus, independent sight feeds. Fig. 539, are sometimes 
used in connection with a central reservoir. Such an installation is 
shown diagrammatically in Fig. 540. A condenser pipe leading from 
the steam main enters the bottom of the reservoir and the condensed 
steam fills up the reservoir as fast as the oil is fed out. The principle 
is the same as that of the simple hydrostatic lubricator. Oil is fre- 
quently injected by mechanical means under a steady pressure gen- 



LUBRICANTS AND LUBRICATION 



797 



erated and governed independently of the steam. Two systems are 
in common use, direct mechanical pump pressure and air pressure. 



(r^ 


1 




1 


^ Steam Main 

H 






s 


T'' 


'( )^ 


"V^ . 


s 




c 


^ . 


' 




( 


>JI 




A 




Cylinder 
Reservoirs 


To Other Engines 
1 




' /-s /^ 









Fig. 540. Central Hydrostatic Lubricator. 

392. Forced-feed Cylinder Lubrication. — Fig. 541 illustrates the 
'^Rochester" simple feed automatic lubricating pump, which takes the 
oil by gravity from the reservoir through a sight-feed glass and forces 
it through a small pipe to the steam supply pipe. The pump entirely 




Fig. 54L Rochester Forced-feed Lubricator. 

obviates the trouble due to intermittent feeding and, being directly 
driven from the engine, runs at constant speed. The feed is uniform 
and independent of the pressure pumped against. The rate is deter- 
mined by the length of stroke of the pump piston, which is easily adjusted. 



798 



STEAM POWER PLANT ENGINEERING 



With large engines multi-feed pumps are sometimes used, which force 
oil to the various valves as well as to the steam pipe. Fig. 542 shows 



H.P.Steam Pipe 

L.P. Steam Pipe j^ ^^^ 

To Rod \ . L, 




Fig. 542. Forced-feed Cylinder Lubrication. 

an arrangement of storage tank in connection with pump reservoir to 
avoid the trouble of hand filUng. 

393. Central Systems. — Fig. 543 shows the piping for a large central 
system of cyUnder and engine lubrication. There are two storage 




OilPump 
FRONT ELEVATION 



Fig. 543. Central System for Large Stations. 

tanks on the engine-room floor, one for cylinder oil and the other for 
engine oil, the distributing arrangements being the same in each case. 
The oil is pumped from each tank into a main pipe extending the length 



LUBRICANTS AND LUBRICATION 



799 




TO STtHM 



rEEO REGULATOR 



of the engine room and provided with branches at each point requiring 
lubrication. The oil pumps are ac- 
tuated by steam and are of the 
duplex direct-acting type, provided 
with automatic governors which 
regulate the speed to suit the de- 
mand for oil. The cylinder oil is 
forced through a special sight-feed 
lubricator, Fig. 544, under a pres- 
sure of about 25 pounds in excess 
of the steam pressure. Referring to 
Fig. 544, diaphragm valve D, in the 
bottom of the lubricator, is kept 
closed by the steam pressure ad- 
mitted through pipes B. Thus the I^i^- ^^^^ ^iegrist Sight-feed Lubricator, 
inlet pressure must be greater than that of the steam before the valve 

will open and admit oil to the 
engine. The oil, after enter- 
ing, passes upward through the 
sight-feed glass and downward 
through the hollow arm A to 
the steam pipe. The engine 
oil is forced by the pump to the 
various points under a pressure 
of about 20 pounds. The waste 
oil is caught in suitable recepta- 
cles and, after being filtered, is 
returned to the storage tank by 
a steam pump. This pump is 
connected so that it can supply 
the storage tank either from the 
filter or with fresh oil from a 
large oil tank in the basement. 
By this arrangement all han- 
dling of oil in the engine room is 
done away with. 

Fig. 545 gives a diagram- 
matic outhne of the oiling sys- 
tem for a vertical Curtis steam 
turbine. A tank, of sufficient 
capacity to contain all the oil 
and fitted with suitable straining devices and a cooHng coil, is located 




To Spring Equalizer 
or Accumulatoc 



Fig. 545. Arrangement of Oiling System for 
Vertical Curtis Turbine. 



800 



STEAM POWER PLANT ENGINEERING 



at a level low enough to receive oil by gravity from all points lubri- 
cated. A pump draws oil from this tank and delivers it at a pres- 
sure about 25 per cent higher than that required to sustain the weight 
of the turbine in the step bearing. A spiral duct baffle connects the 
source of pressure to the step bearing and serves to regulate the oil 
supply to the lower end of the shaft. This source of pressure is 
also connected through a reducing valve to the upper oihng system 
of the machine, in which a pressure of about 60 pounds to the 
square inch is maintained. This system, which includes a storage 
tank partly filled with compressed air, operates the hydraulic governor 
mechanism and supplies oil to the upper bearings. DeHvery of oil 
to these bearings is regulated by adjustable baffles designed to offer 



r^-'^Th 



Gear Lubrication 



Gauge 
Feed Pipe to 
Water Cooled Lining 

K Globe Valve 
Drilled Seat 




Return from Gear 
Casing to Oil TankX ' 



•^ 



1^ Globe Valve 

/' Drilled Seat 
J^ "Gauge Cock 



Tank 




Fig, 546. Diagram of Oil Piping for Curtis Horizontal Turbine. 



resistance to the oil flow without forcing the oil to pass through any 
very small opening which might easily become clogged. A relief valve 
is provided to prevent the pressure in the upper part of the oiling 
system from rising above a desirable limit. Drain pipes from the 
upper bearings and from the hydraulic cylinder and relief valve all 
discharge into a common chamber, in which the streams are visible, 
so that the oil distribution can always be easily observed. At some 
point in the high-pressure system adjacent to the pump it is desirable 
to install a device to equalize the delivery of oil from the pump, as is 
done by the air chamber commonly used with pumps designed for 
low pressure. A small spring accumulator is furnished for this purpose, 
except in cases where weighted storage accumulators are used. In 
large stations where several machines are installed, a storage accumu- 
lator is desirable and can be arranged advantageously so that it will 
normally remain full, but will discharge if pressure fails, and in doing 
so will start auxiliary pumping apparatus. 



LUBRICANTS AND LUBRICATION 



801 



Gaufe't 



All modern steam turbines are equipped with forced feed lubrica- 
tors. The oil pumps are either independently driven or geared to 
the turbine shaft. The different systems employed are described in 
paragraphs 207-213. 

394. OU Filters. — After oil has been applied to machinery its lubri- 
cating properties become impaired on account of (1) contamination 
with anti-lubricating material, such as dust, metallic particles from 
wear, gum, acid, and resin; and (2) exposure to heat and the atmos- 
phere which drives off part of the more volatile constituents and de- 
creases the fluidity of the oil. 

In many small plants no attempt is made to reclaim oil that has 
once been used, since the quantity is so small that the cost and trouble 
involved would more than 
offset the gain. AVhere large 
quantities of oil are used, 
considerable saving may be 
effected by using it over and 
over again. To render the 
oil fit for reuse it must be 
thoroughly purified. The 
anti-lubricating matter is re- 
moved by precipitation and 
filtration. 

Fig. 547 shows a section 
through a ''White Star" oil 
filter and purifier. The ap- 
paratus consists of a cylin- 
drical sheet-iron vessel di- 

\'ided into two compartments by a vertical partition. These two 
compartments are connected near the top by valve B. The smaller 
chamber is provided with a funnel A and a steam coil for heating the 
contents. The large chamber contains a cylindrical wire screen cov- 
ered with several folds of filtering cloth. Impure oil is poured into 
funnel A, the upper part of which is provided with a removable sieve 
or strainer, and is discharged below the surface of the water through 
holes in the foot of the tube. The thin streams of oil rise vertically 
to the surface of the water and the heavy particles of grit and dirt 
gravitate to the bottom. The steam coil heats the oil and water 
and facilitates precipitation of the solid matter by thinning out the 
streams of oil. When the oil in the smaller chamber reaches the level 
of valve B it flows into the filter bag, which removes the remaining 
impurities and permits the purified products to flow into the large 




Fig. 547. White Star Oil Filter. 



802 



STEAM POWER PLANT ENGINEERING 



compartments from which it may be drawn at will. All parts are 
accessible and readily removed for cleaning purposes. The accumu- 
lated sediment in the bottom of the small chamber is discharged to 
waste at intervals by means of a suitable drain. When the filter cloth 
is to be removed, valve B is closed and the wire cylinder is disconnected 
and lifted out. Any oil remaining in the filter is returned to funnel A . 
The filter cloth is held against the screen by cords and hence is readily 
removed. 

Fig. 548 shows a section through a Turner oil filter, illustrating the 
type of filter usually installed in large stations where continuous fil- 
tration is desired. This apparatus consists of a rectangular tank 
divided into four compartments. The returns from the lubricating 



Perforated Plato 

Filtering Material 

Perforated Plat© 



Perforated Plato 

FilteringMaterial 



Perforated Plate 
Water 




Steam Coils 



SECTION 1 SECTION 2 SECTION 3 SECTION 4 

Fig. 548. Turner Oil Filter. 



system flow into section 1 through a screened funnel and discharge 
into the water space at the bottom of the compartment. The oil rises 
through the water, passes, under pressure of the head in the funnel, 
through a layer of filtering material resting on a perforated plate, and 
collects in an inverted cone. Through perforations around the top of 
the cone it passes into a dirt chamber, where most of the heavy im- 
purities are deposited, and then, still rising, passes through another 
perforated plate and more filtering material. The partially cleaned oil, 
which issues, overflows into the second compartment and thence into 
the third, the same cycle of operations being repeated in these two. 
The overflow from the third compartment descends through a final 
filter in the fourth compartment and collects at the bottom, from which 
it is withdrawn by the oil pump. 



p 



LUBRICANTS AND LUBRICATION 803 

Cylinder Lubrieation: Power, Apr. 11, 1916, p. 519, Feb. 15, 1910; Jour. A.S.M.E., 
Feb. and May, 1910. 

Miscellaneous. — Measurement of Durability of Lubricants: Trans, A.S.M.E., 
11-1013. Valuation of Lubricant by Consumer: Trans. A.S.M.E., 6-437. Suit- 
ability of Lubricants: Power, Nov., 1906, p. 673. Oil Required for Lubricators: 
Elec. World, May 5, 1906, p. 934. Gumming Tests: Jour. Am. Chem. Soc, April, 

1902, p. 467. Valuation of Lubricants: Jour. Soc. Chera. Ind., April 15, 1905, 
p. 315. 

Lubrication, General: Prac. Engr., Oct. 1, 1916, p. 833; Power, Sept. 12, 1911, 
p. 396; Sibley Jour., June, 1916, p. 277. 

Oil Purification: Elec. World, Dec. 1, 1906, p. 1053. 

Economy in Lubrication of Machinery: Trans. A.S.M.E., 4-315. Theory of 
Finance of Lubrication: Trans. A.S.M.E., 6-437. 

Experiments, Formulas, and Constants for Lubrication of Bearings: Am. Mach., 

1903, pp. 1281, 1316, 1350. 

Lubricators and Lubricants: Power, Sept. 21, 1909, p. 486, Feb. 22, 1910, p. 347. 
Selection of an Oil for Lubrication: Power, July 27, 1909, p. 137. 
Lubrication with Oils, and with Colloidal Graphite: Jour. Industrial and Engineer- 
ing Chemistry, Vol. 5, No. 9, Sept., 1913. 

Tests of Used Oil: Prac. Engr., Apr. 15, 1914, p. 469. 

Laws of Lubrication of Journal Bearings: Trans. A.S.M.E., 37-1915, p. 534. 



CHAPTER XVII 



TESTING AND MEASURING APPARATUS 



395. General. — The importance of maintaining a system of records 
is discussed in paragraph 419. The various items which may be re- 
corded and the instruments and appHances used in this connection are 
outHned in the accompanying chart. In large stations a full comple- 
ment of indicating, recording, and integrating instruments may prove 
to be a good investment if intelHgently and closely studied by the 
operating engineer with a view to locating and ehminating unnecessary 
losses. The instruments should be inspected and calibrated at inter- 
vals, since many of them are delicately constructed and are apt to 
become inaccurate after a few months' service. Steam gauges, ther- 
mometers, and pyrometers, and particularly piston 
water meters are subject to appreciable error after 
lU fJ- — I considerable use. Voltmeters, ammeters, and other 

switchboard instruments are easily deranged, es- 
pecially when subjected to continuous vibration or 
to high temperature. 

396. Weighing the Fuel. — In most small plants 
the delivery tickets of the coal dealer are depended 
upon for the weight of coal used, no attempt be- 
ing made to determine the evaporative value, and 
the economy of the plant is judged by the size of 
the coal bill. In such cases a considerable saving 
may be effected by keeping a daily record cover- 
ing at least the coal and water consumption. The 
coal can be conveniently weighed on ordinary 
•v.;— .:<,•;■. ^ platform scales. In a number of large stations 
£MS0Mk^ the weight of coal is determined by suspended 
weighing hoppers, which may be stationary, as in 
Fig. 141, or mounted on a traveling truck, as in Fig. 
142. The scales of such devices are made indicat- 
ing, autographic, integrating, or a combination of the three, the latter 
costing but little more than the simple indicating or recording devices. 
A simple and inexpensive coal meter recently brought out is illus- 
trated in Fig. 549. It consists essentially of a helical vane placed in a 
cyUndrical conduit. The movement of the coal causes the vane to 

804 



Fig. 549. Coal Meter. 



TESTING AND MEASURING APPARATUS 



805 



TESTING AND MEASURING APPARATUS. 



i 



Weights 



Pressures 



Tempera- 
tures 



I 



Power , 



Flue gas 
analysis 



Moisture. 



Fuel analysis 



Steam Plant. 



rFuel 



Platform scales, indicating and autographic. 
Suspension hoppers, indicating and auto- 
graphic. 
Coal meters, integrating. 
Platform scales and tanks, 
r Piston .. ) 
Rotary. . > Integrating. 
Water meters . . -( Disk . . . . ) 

Venturi, indicating and 
autographic. 

Weirs and volume displacement meters. 
c, ( Weighing condensed steam. 

•Steam ) ^ ^ , Dirpct 

^Steam meters.. I j^Xe^^_ 

Bourdon gauge, indicating and autographic. 
Manometers, mercurial, indicating. 
Manometers — mercurial, indicating, and 

autographic. 
Manometers — water, indicating, and auto- 
graphic. 
.Diaphragms, indicating and autographic. 
( Mercurial thermometers, indicating. 
< Expansion thermometers, indicating and 
( autographic. 
Expansion thermometers, indicating and 

autographic. 
Resistance thermometers, indicating and 

autographic. 
Thermo-electric thermometers, indicating 

and autographic. 
Optical pyrometer, indicating and auto- 
graphic. 
Platinum or clay ball pyrometer. 
, , ( Indicators, hand manipulated. 

^ "I Indicators, continuous autographic. 

{Rope brake. 
Prony brake. 
Absorption dynamometers. 
Electric generator. 
Orsat apparatus. 
Hay's recorder. 

Westover recorder, autographic, 
Uehling gas composimeter, autographic. 
In air Hygrometer, indicating and autographic. 

In steam Calorimeters . . 

[ Mahler bomb. 
Coal calorimeters. . . . ] Thompson. 

Parr. 
Gas calorimeter Junker. 




Indicat 



Separating. 
Throttling. 



Voltage 

Current 

Output 

Power factor. , 
Frequency. . . . 
Synchronism. . 



Electrical Plant. 

Voltmeters, A. C. and D. C, indicating and autographic. 
.Ammeters, A. C. and D. C, indicating and autographic. 
Wattmeters, A. C. and D. C., integrating and autographic. 
Power factor meters, A. C. only, indicating and autographic. 
Frequency meter, A. C. only, indicating. 
Synchronizers, A. C. only, indicating. 



806 STEAM POWER PLANT ENGINEERING 

rotate and the number of revolutions is a measure of the weight of fuel 
passing. For hard coal of uniform size the meter gives consistent 
results agreeing within two per cent of scale weight, but with bitumi- 
nous coal the results are somewhat erratic and particularly so with 
lumps of varying size. (For a detailed description of the device, see 
Prac. Engr., U. S., Apr. 15, 1912, p. 438.) With certain types of me- 
chanical stokers it is possible to approximate the rate at which fuel 
is fed into the furnace by registering the speed of the stoker engine. 
In the new River Station of the Buffalo General Electric Co. ''Electric 
stoker tachometers" are used for this purpose. 

397. Measurement of Feed Water. — The quantity of water fed to 
the boiler may be determined by 

1. Actual weighing. 

2. Measurement of volume displacement. 

3. Measurements by weirs and orifices. 

4. Measurement by determining the velocity of flow in the feed pipe. 
Some of these methods necessitate measurement on the suction side 

of the pump; others are appUcable to either suction or pressure. The 
former, as a class, are the more accurate but involve bulky apparatus. 
The choice for any given case depends upon the quantity of liquid to 
be measured, the degree of accuracy required, space requirements, and 
first cost. 

398. Actual Weighing of Feed Water. — The most accurate means of 
measurement is by the use of two or more tanks resting upon scales, 
arranged to be filled and emptied alternately. This method is limited 
to comparatively small quantities because of the great bulk of appa- 
ratus involved and is seldom used for continuous service. It is com- 
monly employed in conducting special tests of short duration and for 
calibration purposes. For regular boiler service it involves considerably 
more time than is ordinarily at the disposal of the fireman and engineer. 
For temperatures above 150 deg. fahr., the weighing tanks should be 
covered, since evaporation may cause an appreciable error. See also 
"Rules for Conducting Boiler Trials," A.S.M.E., Code of 1915. 

399. Worthington Weight Determinator. — Fig. 550 shows the gen- 
eral details of the Worthington weight determinator, illustrating a 
commercial means of continuously measuring and recording the weight 
of water fed to the boiler. The apparatus consists primarily of two 
tanks of equal size, A and B, each mounted on knife edges K and 
equipped at one end with a siphon S and at the other end with counter- 
weight W. The hquid to be measured flows through inlet pipe P and 
along deflector D into either tank. Each tank remains in a horizontal 
position until the weight of liquid overcomes the counterweight when 



TESTING AND MEASURING APPARATUS 



807 



it tilts into the position shown by the dotted Hnes. Discharge now 
takes place through siphon S until the liquid reaches a certain level at 
which point the tank tilts back to its original position and the siphon 




SECTION X-Y 

Fig. 550. Worthington Water Weigher. 

continues its action until the vessel is emptied. The tanks operate 
alternately, one filling while the other is discharging. Since each tilt 
represents a definite weight of liquid irrespective of variations in volume 
due to specific gravity or changes in 
temperature, the number of tilts as 
recorded by counter C is a correct 
measure of the weight discharged. 
This apparatus operates at atmos- 
pheric pressure and is arranged to 
discharge into a storage tank from 
which the feed pump takes its supply. 
400. Kennicott Water Weigher. — 
This apparatus is used in many boiler 
houses and seems to give universal 
satisfaction. It consists of a cyhndri- 
cal shell *S, Fig. 551, the lower part of 
which is divided into two measuring 
compartments A and B, each fitted 
with a siphon for discharge and a float 
F for actuating the tripping mecha- 
nism. Tripping box E is divided into 
two sections which alternately fill with water and serves the double 
purpose of furnishing a sufficient quantity of water to start the si- 
phons and to shift the supply from one compartment to the other. 
This tripping box is balanced on knife edges and is mounted directly 




DisdiaFg* 



Fig. 551. Kennicott Water 
Weigher. 



808 



STEAM POWER PLANT ENGINEERING 



above the measuring compartments. Water enters the inlet and passes 
to the tripping box where a small portion is intercepted, the re- 
mainder passing directly to the measuring compartment below. When 
this compartment is nearly filled the float tilts the tripping box, dis- 
charges its contents into the compartment, and starts the siphon. A 
counter registers each double charge. This apparatus discharges at 
atmospheric pressure, though with slight modification it may be in- 
stalled on the pressure side of the pump. Kennicott water weighers 
are constructed in various sizes ranging from a capacity of 750 to one 
miUion pounds per hour and are guaranteed by the manufacturers to 
record the correct weight of water within one-half of one per cent of 
scale weight at any given temperature. Calibration for different tem- 
peratures is necessary since the apparatus is actuated by volume dis- 
placement. For example, the weight of one cubic foot of water at 
60 deg. fahr. is 62.37 pounds and at 210 deg. fahr. it is 59.88, a differ- 
ence of 2.49 pounds. Hence, if the device is calibrated to read correctly 
at 60 degrees it would be in error 4 per cent if used to measure water 
at 210 deg. fahr. 

401. Willcox Water Weigher. — Another successful volume displace- 
ment meter is illustrated in Fig. 552. The device consists of a cylin- 
drical tank divided into an up- 
per and lower compartment by 
a horizontal partition. The 
water enters the upper com- 
partment, passes to the lower, 
in which its volume is meas- 
ured, and then out through the 
U-shaped discharge pipe. The 
operation, beginning with both 
compartments empty, is as fol- 
lows: Water enters the upper 
compartment through the inlet 
pipe and rises to the top of the 
standpipe. (The latter is open 
at the top and bottom and is 
rigidly connected to the bell 
float, but when in its lowest 




Discharge Pipe 



Willcox Water Weigher. 



position it is held against its seat by weight of the bell float.) Further 
admission of water causes it to overflow into and through the stand- 
pipe into the lower compartment. The water, rising in the lower com- 
partment, seals the lower edge of the bell float and entraps a volume 
of air under the bell. Further rise compresses the air under the float, 



TESTING AND MEASURING APPARATUS 



809 



in leg C of the discharge pipe and in leg A of the trip pipe AB. This 
compression causes the float to rise to its highest position and raises the 
standpipe from its seat, permitting the water in the upper chamber to 
pour into the lower vessel. Compression of air continues until the 
pressure becomes great enough to break the seal in the trip pipe. This 
action immediately reduces the pressure below the float, permits the 
latter to descend, sealing the upper chamber against further discharge, 
and allows the water in the lower compartment to siphon out through 
the discharge pipe. The number of discharges is recorded mechanically. 




Fig. 553. A Typical Piston Water Meter. (Worthington.) 

402. Weir Measuring Devices. — Feed water heaters or specially 
designed tanks fitted with V-shaped, cycloidal, or trapezoidal weir 
notches offer a simple means of measuring the rate of flow. The 
chamber is divided into vertical compartments arranged so that one 
may discharge through a cahbrated weir notch into the other. The 
height of water above the bottom of the notch is a direct measure 
of the volume flowing. The height may be noted in an ordinary gauge 
glass or it may be transferred through a suitable float mechanism to an 
outside indicator. Commercial weir measuring devices are usually 
provided with autographic and integrating attachments for recording 
the rate of flow and for totahng the weight of water passing through 
the device. For the theory of weir notches, orifices, and nozzles consult 
''Experimental Engineering," Carpenter and Diederichs, 1911, Chapter 
XII. See also. Trans. A.S.M.E., 1915. 

Weir Meters for the Power Plant: Power, May 1, 1917, p. 582. 

403. Pressure Water Meters. — There are a number of reliable water 
meters on the market for hot or cold water which may be placed on the 
pressure side of the feed pump. Among them may be mentioned the 
Hersey, Crown, Nash, and Worthington. They are all ])ascd on vol- 
ume displacement and consequently require correction for different 



810 



STEAM POWER PLANT ENGINEERING 



temperatures if graduated to read in pounds. They are compact, 
comparatively inexpensive, and require considerably less space than the 
tank weighers of the Kennicott and Willcox type but are open to the 
objection that no particular provision is made against leakage and after 
considerable use they are subject to serious error. In many plants 
where meters of this type are installed the meter is by-passed and 
operated only for short periods. For continuous service meters of the 
tank-weighing or Venturi type are recommended. Fig. 554 illustrates 




Fig. 554. A Typical Disk Water Meter. (Nash.) 

the piston type of pressure meter, in which reciprocating pistons are 
displaced by a definite volume of water; Fig. 425, the rotary type, de- 
pending upon the displacement of rotary impellers; Fig. 555, the disk 
type, in which impellers are given a combined rotating and tilting 
motion. The capacities of pressure meters range approximately as 
follows : 

Size of meter (pipe size) f , 

Maximum capacity, cubic feet per minute : 

Rotary or disk meters 1, 

Piston meters 1^, 

404. Venturi Meter. — The Venturi tube with indicating, auto- 
graphic, and integrating mechanism, as constructed by the Builder's 
Iron Foundry of Providence, R. I., is one of the most satisfactory 
methods of measuring feed water under pressure. The total absence 
of working prrts in the meter proper insures continuity of operation 
and freedom from wear, and the fact that the recording mechanism 
may be placed at a considerable distance from the meter is a great 



h 


3 

4J 


1, 


U, 


2, 


3, 


4, 


6 


2, 


4, 


8, 


12, 


20, 


36, 


72, 


120 


3. 


5, 


6, 


8, 


23, 


60, 


120 





TESTING AND MEASURING APPARATUS 



811 



advantage. The Venturi tube, Fig. odd, is essentially the same in 
principle as an orifice placed in the pipe. The pressure difference H 
between A in the ''upstream" portion of the tube and B at the ''throat" 
is a measure of the velocity through the throat. The loss of head due 



Pipes to Manometer 



LU4H1^^ 



Outlet 




Fig. 555. Venturi Tube with Indicating Manometer. 

to friction is negligible and the velocity may be calculated, within an 
error of 2 per cent, from the following modification of Bernouilli's 
theorem : 



Vt 



V2^, 



(299) 



VFJ - Ft' 
in which 

Vt = velocity at the throat, feet per second, 
Fu = area of the upstream section, square feet, 
Ft = area of the throat, square feet, 
H = pressure difference, feet of water. 

For accurate work the tube requires calibration. Once calibrated the 
error in weight readings for a given temperature should not exceed one 
per cent for capacities within the working range of the manometer. 
For very low throat velocities the error may be considerable because 
of the shght pressure difference between the points A and B. In situa- 
tions where there are periods of very low and very high rates of flow, 
as in connection with combined heating and lighting plants, it is cus- 
tomary to install a small tube for the light loads and a large tube for 
the heavy loads, the same indicating mechanism being used in each 
case. The equipment illustrated in Fig. 555 is purely indicating and 
readings must be taken at frequent intervals in order to obtain the 
total flow for a given period. Where the size of the plant warrants the 
outlay the combined indicating, integrating, and recoidilig instrument 
is often installed. With this device the instantaneous rate of flow is 
indicated by a pointer and dial, the variation in rate of flow for any 



812 



STEAM POWER PLANT ENGINEERING 



given period is recorded on a clock-driven chart, and the total weight 
flowing is registered on a counter. (For a detailed description of this 
mechanism see Power, Jan. 23, 1912, p. 102.) Tests made at Armour 
Institute of Technology on a carefully calibrated tube and recorder 
with feed water at 210 deg. fahr. and constant rate of flow gave chart 
and counter readings agreeing substantially with scale weights; for 
irregular and fluctuating flow, as when feeding the boilers, the average 
error was about two per cent. 

405. Orifice Measurements. — The appropriation of the great majority 
of small steam power plants does not permit of the installation of 
tank meters, Venturi meters, or other forms of reliable commercial 
appliances for measuring the weight of water fed to the boilers. For 




-1 Solid Iron Bac 
• 1 Iron Gage Cock 

Fig. 556. Simple Indicating Water Meter, Orifice Type. 

use in such cases an inexpensive and fairly accurate indicating meter 
may. be constructed of ordinary pipe fittings, as illustrated in Fig. 556. 
A thin metal diaphragm with circular orifice is inserted on the pres- 
sure side of the feed pump and the pressure drop across the orifice is 
measured by incHned mercury manometer. The height of mercury 
h is an indication of the rate of flow. By calibrating the manometer 
against tank measurements the readings of the mercury column may 
be graduated to read directly in pounds per hour. If means are not 
available for calibration purposes the weight of discharge may be 
approximated from the formula 

W = 1120 a Vm, (300) 

in which 

W = weight flowing, pounds per hour, 
a = area of the orifice, square inches, 
h = vertical height of mercury column, inches, 
d = density of the water, pounds per cubic foot. 



TESTING AND MEASURING APPARATUS 818 

For a fairly continuous flow and pressure drop corresponding to three 
inches of mercury or more this simple device gives results agreeing 
within four per cent of tank weights, but for widely fluctuating flow 
and small pressure drops the error may be considerably more. 

For application of the Pitot tube for water measurements consult 
accompanjdng bibliography. 

The Pitot Tube for Water Measurements: Trans. A.S.M.E., Vol. 30, 1908, p. 3ol, 
Vol. 25, 1904, p. 184, Vol. 22, 1901, p. 284. 

The Pitometer: Proc. Am. Wks. Asso., 1907, p. 136; Jour. Frank. Inst., D(;c., 
1907, p. 425. 

406. Measurement of Steam. — The quantity of steam passing 
through any device may be determined by (1) condensing and weighing 
the steam after it has passed through the apparatus and by (2) meas- 
uring the flow by means of steam meters before it enters. The first 
necessitates the use of surface condensers, and consequently has a 
limited field of application, whereas the latter may be used in both 
condensing and non-condensing service. 

40'7. Weighing Condensed Steam. — The weight of condensed steam 
may be obtained by any of the devices used in connection with feed 
water measurements but such measurements are seldom made except 
for test purposes because of the expense or labor involved. The Wheeler 
Condenser and Engineering Company's ''indicating hot well" offers a 
practical and simple solution of continuously measuring the condensed 
steam. The hot well is attached to the bottom of the condenser 
chamber in the usual way and differs from the ordinary hot well only 
in the addition of a vertical partition. This partition divides the hot 
well chamber into two compartments. Condensation from the con- 
denser drains into one of these compartments and flows to the other 
through a calibrated orifice. The height of water above the orifice as 
shown in the gauge glass is an indication of the weight of condensation 
flowing. By means of suitable attachments the readings may be 
automatically recorded and totaled. The manufacturers guarantee an 
accuracy within 2 per cent of scale weight for readings over the whole 
range. 

408. Steam Meters. — The weight of fluid flowing through an open- 
ing may be calculated by the equation 

m which ^ ^ 

W = weight in pounds per second, 
A = cross-sectional area in square feet, 
y = density of the fluid, pounds per cubic foot, 
V = velocity of flow, feet per second. 



814 



STEAM POWER PLANT ENGINEERING 



All steam meters for indicating or recording the weight of steam 
flowing through a pipe are based upon the law expressed in equation 
(301). Thus, for steam of constant density the opening through which 
it flows may be made constant and the variation in velocity will be an 
indication of the rate of discharge; or the velocity may be held constant 
and a variation in the amount of opening will be an indication of the 
weight discharged. Unfortunately, the density of steam is seldom 
constant under commercial conditions and herein hes the inherent 
defect of all steam meters which depend for their operation upon a 
variation in the area of efflux or a variation in velocity. The density 
of steam is a function of its pressure and quality and any variation in 
either will affect the weight of discharge as determined from equation 
(301). Pressure variations may be automatically compensated for, but 
corrections for quality must be made in each specific case. 



CLASSIFICATION OF STEAM METERS. 



Indirect 



Direct 







Impeller 




Pitot tube 


Water 

manometer 


Velocity 




Mercury 
manometer 




, Current 


Impeller 




Floating 
valve 


Mechanical 
control 


Throttling , 


Stationary 
disk 

Venturi 
, tube 


Mercury 

manometer 
Bourdon 

manometer 

Mercury 
manometer 



Lindenheim (1896)* 
Gebhardt (1908)t 
Burnham (1905)t 

. Gebhardt (1910)t 

( General 

] Electric (1910)*tt 

( RepubUc 1916*tt 
Holly (1877)* 
St. Johns (1893)tt 
Gehre (1896)tt 
Baeyer (1902)tt 
Bendemen (1902)t$ 
Sargent (1908)t 
LindmarkfJ 
Gehre-Hallwachs 
(1907-1910)*tJ 
Sarco (1910)*tJ 
Bailey (1910)*tt 
Eckardts (1903) ft 



( Parenty (1886) ft 

] Builders' Iron 

( Foundry (1910)t$ 

* Integrating. f Indicating. J Autographic. 

The different means adopted for transmitting this area and velocity 
variation to the indicating or recording devices overlap to such an 
extent as to render a classification of steam meters very unsatisfactory. 
The accompanying chart is offered as a guide in grouping the most 
commonly known devices. From this chart it will be seen that all 
meters may be grouped into general classes, direct and indirect. The 
direct meter is an integral part of the piping and the entire mass of 
fluid to be measured passes through the apparatus. It is not portable 



TESTING AND MEASURING APPARATUS 



S15 



and cannot be readily applied to pipes of different sizes. In the in- 
direct meter only a small part of the fluid to be measured is directed 
through the apparatus and the pipe line need not be disconnected for 
its installation. One instrument suitably calibrated may answer for 
any size of pipe. 

The average high-grade steam meter is a reliable and accurate means 
of measuring the flow of steam in straight lengths of pipes, provided 
the flow is continuous or that the change in the rate of flow is gradual 
and the pressure and quality are practically constant. For interrupted 
or intermittent flow and for sudden variations in pressure or quality, the 
results are not reliable and may be considerably in error. The accu- 
racy of all meters, provided they have been correctly calibrated and 
adjusted, depends largely upon the degree of refinement in reading the 
indicators and in integrating the charts. The commercial failure of 
many steam meters is due to the fact that they are not cared for or 
operated in strict accordance with the principles of design. 

Only a few of the best-known meters will be described here. For a 
detailed discussion of the various types of steam meters see the author's 
paper "Various Types of Steam Meters," Power, Feb. 6 and 13, 1912. 




Figs. 557, 558, 559. Principles of the "Gebhardt" Indicating Steam Meters. 

"Gebhardt" Steam Meters. — Figs. 557 to 560 illustrate various 
forms of indicating steam meters designed and tested at the Armour 
Institute of Technology, which are based on the principles of the 
Pitot tube. Referring to Fig. 557, A and C are two ordinary gauge 
cocks and G is a common gauge glass, C being connected with the static 
nozzle S and A with the dynamic tube D. The height of water H is 
proportional to the square of the velocity of steam flowing through 
pipe P and automatically adjusts itself to the variations in velocity; 
thus, for decreasing velocities, the water in glass G discharges through 
D until the water column H balances the velocity pressure in pipe P, 



816 



STEAM POWER PLANT ENGINEERING 



and for increasing velocities, condensation from the upper part of the 
instrument accumulates and the water column H rises until a balance 
is effected for the higher velocities. 




Fig. 560. Commercial Form of "Gebhardt" Steam Meter. 

The relation between the height of the water column and the velocity 
of the steam in the main pipe at the entrance to the dynamic tube may 
be determined from the well-known equation 



V = c V2 gh, 



(302) 



in which 

V = maximum velocity of flow, feet per second, 
c = coefficient determined by experiment, 

h = height of a column of steam equal in weight to the water 
column H. 



The equation may be expressed 



V 



k^h'I 






(302a) 



in which 

K = coefficient determined by experiment, 

H = height of water column in inches, 

dy, = density of water in gauge glass, pounds per cubic foot, 

ds = density of steam in the main pipe. 

Because of the labor of determining the relationship between the 
mean and the maximum velocity for various conditions of flow and 
different pipe diameters it is more satisfactory to calibrate the gauge, 
by actual experiment, to read directly in pounds per hour. 



TESTING AND MEASURING APPARATUS 



817 



This simple device in connection with a caUbrated scale gives readings 
within 5 per cent of condenser measurements for continuous flow and 
constant pressure and quality of steam (for velocity pressures corre- 
sponding to IJ inch of water or more). For a considerable variation in 
pressure and quality or for marked changes in rate of flow the instru- 
ment is not rehable. Its sensitiveness is greater at high velocities, 
since the height of water column in the gauge glass increases with the 
square of the velocity of the steam in the main pipe. For interrupted 
.low, as when connected to a high-speed engine, the water column may 
le made to closely approximate the mean velocity of suitably throttling 
the gauge cocks. 

Fig. 558 shows application of the same principle with only one con- 
nection to the main pipe. Under favorable conditions the commercial 
meter (Fig. 560) gives readings within 2 per cent of condenser weights 
for velocity pressures corresponding to 1 inch of water or more. Fig. 
559 shows another form which may be placed below or above the point 
in the main pipe at which the Pitot tubes are placed. The operation 
is as follows: Velocity pressure is transmitted through tube D and 
opening 0, into the body of the chamber M. This pressure, acting on 
the surface of the condensed steam in the chamber, forces the water 
into the glass W until a balance is effected. Condensation is dis- 
charged continuously through pipe P and the water seal U of the main 
pipe. Tests of this meter have given re- 
sults agreeing within 2 per cent of con- 
denser measurements for continuous flow 
for all velocities ranging from the equiva- 
lent of a 1-inch to a 10-inch water column. 
No provision is made for automatic correc- 
tion of pressure and quality variation in 
any of these devices. (For the theory and 
results of tests of the Pitot type of steam 
meter see author's paper "The Pitot Tube 
as a ^team Meter," Trans. A.S.M.E., Vol. 
31, p. 603.) 

G'E. Flow Meters. — All G-E. flow me- 
ters, with the exception of the '^orifice 
tube" type for small pipe sizes, depend for their operation upon the 
displacement of a mercury column by the differential pressure action 
of a modified Pitot tube. The basic principle of operation is illus- 
trated in Fig. 561: S is the static opening and D the dynamic open- 
ing; U is an ordinary U-tube manometer partially filled with mercury. 
When there is no flow the surface of the mercury in columns iV and TF 




Fig. 561. Pitot Tube with 
Mercury Manometer. 



818 



STEAM POWER PLANT ENGINEERING 



will be on the same level and the upper portion will be filled with con- 
densed steam. When there is a flow, the mercury will be depressed as 
indicated and the difference H will be a measure of the velocity of 
flow at the point in the pipe where the dynamic tube is placed. This 
velocity may be expressed by the equation 



V = K^H'f, 



(302b) 



in which 

dm = density of mercury in lb. per cu. ft. 

Other notations as in equation (302a). 

A comparison of equations (302a) and (302b) will show that the mer- 
cury manometer is less sensitive than the water manometer by an amount 
equivalent to dm -^ dw, or approximately 13.6. The variable heights 
of the water column above the mercury is usually included in the value 
of the coefficient K. 

In all G-E. meters (the "orifice-tube" type excepted) the Pitot tube 
is given the form of a ''nozzle plug" as shown in Fig. 562: TT are the 




Plugged 



Fig. 562. Nozzle Plug; G-E. Steam Meter. 

static openings or ''trailing set" and LL the dynamic openings or ''lead- 
ing set." The plug is screwed into the pipe with the "leading set" 
directly facing the current and the connections to the manometer are 
made through the openings T and L. The manometer for the " portable 
indicating" or laboratory device is shown in Fig. 563. Adjustments for 
variations in pressure, quality, and pipe diameter are made by setting 
the chart cylinder C in accordance with the auxihary scale attached 
to the instrument. The meter may be used to measure flow under 
normal conditions in any number of different pipe lines. It is only 
necessary to provide the pipes with the proper size and kind of nozzle 
plug or pipe reducer to which the meter can be connected. 

Fig. 564 shows a section through the G-E. indicating flow meter 
which differs from the simple portable device in that the movement 



TESTING AND MEASURING APPARATUS 



819 



of the mercury column is magnified by suitable mechanism. A small 
float resting on the top of the mercury in one leg of the U-tube is at- 
tached to a silk cord passing over a pulley; this cord is kept taut by 
a counterbalance weight acting in the opposite direction. The shaft 
on which the pulley is mounted carries a small horseshoe magnet with 
its pole faces near and parallel to the inside surface of a copper plug fas- 
tened to the body of the meter. A small magnet is mounted on pivot 
bearings in such a manner that its poles are near and parallel to the out- 
side surface of the copper plug, and its axis of rotation in Hne with the 



To "Trailing 
Sef ' \ 

r 




Target 



To ' 'Leading Sef 



Adjustment^iHli-i-P 
for Quality ' 



Adjustment for 
Pipe Diameter 
I Adjustment? 



Adjustment for for Pressure 
Height of Chart 

Fig. 563. General Principles of 
the G-E. Indicating-flow Meter. 




/-To Nozzle 

ri=\ Plug 



Fig. 564. Section through G-E. Steam- 
flow Meter. 



shaft carrying the magnet inside the case. The indicating needle is 
attached directly to this magnet. By means of the float and cord, the 
pulley carrying the magnet inside the body is rotated in proportion to 
the change of level of the mercuiy. Any motion of this magnet is 
transmitted magnetically to the outside magnet carrying the indicating 
needle. In cases where the velocity is too low to be accurately measured 
with a normal velocity nozzle-plug, pipe reducers, as illustrated in Fig. 
565, are employed. 

The "G-E. Indicating Recording" meter differs from the simple in- 
dicating device just described only in minor detail. The movement 
of the float is transmitted to the indicating needle and recording pen 
through the agency of a rack and pinion in place of the cord and pulley. 



820 



STEAM POWER PLANT ENGINEERING 




Boiler 



The indicating needle is attached directly to the outside magnet but 
the recording pen is actuated by a sector which in turn is rotated by a 
small pinion on the shaft carrjdng the outside magnet. 

The ^'G-E. Indicating Recording, Integrating" 
meter is identical with the indicating-recording 
device with the exception that an integrating 
mechanism is attached to the sector actuating the 
recording pen. 

For pipes 2 inches or less in diameter the nozzle 
plug is replaced by an ''orifice tube," Fig. 566, 
7^=^^^ j which is to all intents and purposes a Venturi tube. 
Fig. 567 gives a diagrammatic outhne of the 
counting mechanism of a European steam meter 
which serves to illustrate the basic principle of the 
G-E. integrating attachment. 7^ is a small friction 
wheel mounted on the pen arm a and connected to gears c and d by the 
small shaft m; P is a clock-driven disk in contact with the friction wheel 
R. As the pen arm moves the wheel R in and out from the center of 
disk P, the speed of the small friction wheel is decreased or increased 
accordingly. The revolutions of R are transmitted to the integrating 
mechanism e so that the total flow may be read directly from the dials. 

H 



Pipe L 



Fig. 565. Reduciii 
Nozzle. 





Fig. 566. G-E. Orifice-tube 
Steam Meter. 



Fig. 567. Counting Mechanisms for 
Steam Meters. 



Since the pen arm or its equivalent in the G-E. meter does not move 
directly proportional to the velocity of the steam it is necessary to cor- 
rect its movement by means of a cam so that this result may be effected. 
Republic Flow Meter. — This meter is of the Pi tot tube and mercury 
manometer type but differs radically from the G-E. devices in the 



TESTING AND MEASURING APPARATUS 



821 



manner of utilizing the displacement of the mercury columns for in- 
dicating, recording, and integrating purposes. 

The principles of operation are illustrated in Fig. 568; Ci, C2, cz are 
electric conductors, of varying length. As the mercury in the static 



Dynanjif Prebaure 



Static Pressure 



ResiEt^Dces 
Couductors 



-Mercury 




Fig. 568. Fundamental Principles of the Republic Flow Meter. 

leg of the manometer rises it makes successive contact with these con- 
ductors. The resistance of each conductor is such that a constant 
electromotive force impressed upon the circuit will cause a current to 
flow through the conductor directly proportional to the flow of steam 
in the pipe. Any suitable ammeter and watthour meter may there- 
fore be used for indicating, recording, and totaling the weight of steam 
flowing through the pipe. 

Fig. 569 shows the general assembly of the Pi tot tubes or ''tube 
holder" as used in the commercial instrument, and Fig. 570 shows a 



% Plugs 



Dynamic Reservoir 
Static Reservoir 



ileservoir Valve 

To Manometer 

Fig. 569. "Tube Holder," Republic Flow Meter. 

section through the meter body. Referring to Fig. 569 it will be seen 
that the dynamic and static elements are plain cyhndrical tubes with 
beveled ends and placed side by side as indicated. This beveling of 
the ends insures the necessary pressure difference for actuating the 




822 



STEAM POWER PLANT ENGINEERING 



manometer. Referring to Fig. 570, the conductors consist of a large 

number of small steel rods of vary- 
ing length, the lower ends of which 
when in contact with the mercury 
form one end of the circuit; and 
the upper ends in series with in- 
dividual resistance coils are con- 
nected to a common terminal post 
to form the other end of the cir- 
cuit. The conductors and resist- 
ances are insulated by means of 
oil which entirely fills the '^ con- 
tact chamber" above the mercury 
and also the annular chamber be- 
tween the meter body and contact 
chamber. This prevents water 
and foreign substances from reach- 
ing the contact rods. A small 
rotary converter (for direct-current 
supply) or a small transformer (for 
alternating-current supply) fur- 
nishes the necessary current under 
a pressure of 40 volts for actuat- 
ing the various measuring instruments. The maximum current demand 
is approximately one ampere. The particular feature of this meter is 




Fig. 



570. Section through Body of the 
Republic Flow Meter, 




Unit No. 6 



Fig. 571. 



Typical Arrangement of Republic Flow Meter in a Six-unit 
Boiler Plant. 



TESTING AND MEASURING APPARATUS 



823 



that the reading dials can be located at any point with respect to the 
meter body and at any distance from the pipe. Fig. 571 shows a dia- 
grammatic arrangement of a typical installation. The Pitot tubes and 
meter bodies are connected in the boiler outlet. The indicators are 
mounted on the boiler fronts and show the rate of evaporation. The 
board located in the office of the chief engineer includes one indicator, 
integrator, and recorder, and is equipped with suitable switches so that 
the performance of any boiler may be observed at any time. 

St. Johns Steam Meter. — In the groups of meters described above 
the indicating and recording mechanism is actuated by the natural 
velocity of the steam. In the St. Johns, Bailey, Gehre-Hallwachs, 
Storrer, Eckardt, and Venturi steam meters the velocity is increased by 
throttling and the pressure drop is utilized in actuating the mechanism. 
The weight of steam flowing through the orifice may be calculated from 
the following modification of equations (301) and (302) : 



(303) 



W = AK Vp, - p2, 
in which 

W = pounds discharged per second, 
A = area of the orifice, square feet, 
K = coefficient determined by experiment and includes the density 

of the steam, y""^\ 

Pi and p2 = pressure on the upper and 

lower side of the orifice, pounds 

per square inch. 

In some of the meters the pressure drop 
Pi — p2 is maintained constant and the vari- 
ation in the area A actuates the indicating 
mechanism, and in others the area is made 
constant and the variation in pressure drop 
operates the mechanism. 

Fig. 572 represents a section through a 
St. Johns steam meter, illustrating the 
throttling type with a floating valve. This 
meter was placed on the market 20 years ago 
and still finds favor with many engineers. It 
records the weight of steam passing through 
the seat of an automatically lifting valve 
which rises and falls as the demand for steam increases or diminishes. 

Referring to the illustration, valve V is weighted so that a pressure 
in space A of 2 pounds greater than in B is necessary to raise the val'^e 
off its seat. This pressure difference is constant for all positions of the 




St. Johns Steam 
Meter, 



824 



STEAM POWER PLANT ENGINEERING 



valve. The plug is tapered so that the rise of the steam pressure is 
directly proportional to the volume of steam flowing through the seat. 
The movement of the valve is transmitted through suitable levers to an 
indicating dial and a recording pen so that the instantaneous and con- 
tinuous rate of flow may be read at a glance. For a given pressure 
and quahty of steam, the indicating dial and chart may be calibrated 
to read the weight of discharge directly, corrections being made for 
variations in pressure and quality. The manufacturers guarantee the 
readings of the chart to be within 2 per cent of condenser measurements 
for a total pressure range of 10 pounds from the mean pressure at which 
the chart is calibrated. 

The chief drawback to this instrument is inherent to all meters of the 
direct type in that they are bulky and the steam line must be taken down 
for the installation. The total hourly flow may be obtained by inte- 
grating the curve. Tests of this meter made by the author were in 
accordance with the guarantee of the manufacturer for continuous flow 
and for moderate changes in the rate of flow. For rapid fluctuations in 
flow the results were not so satisfactory, the greater error lying in the 
difficulty of integrating the curve correctly. 



Pressure Tight 
Bearing 




.Bell 



-Mercury 
Reservoir 



j^^Bell Weight 
Bell Casing 



Fig. 573. Section through Meter Body of a Bailey Fluid Meter. 

Bailey Fluid Meter. — Fig. 573 shows a section through the manom- 
eter body of a Bailey steam meter. An orifice placed in the steam 
fine at a suitable point effects the necessary pressure drop for actuating 
the mechanism. The higher pressure is appfied at Pi and the lower 



TESTING AND MEASURING APPARATUS 825 

pressure at Pi through small tubes or pipes. The interior of the ''bell 
casing" is subjected to pressure P-i and the interior of the mercury 
sealed ''bell" is subjected to the higher pressure Pi. This difference 
in pressure pushes the bell upward and as it rises from the mercury the 
change in the buoyant action of the mercury on the walls of the bell 
balances the force due to the pressure difference. By varying the 
area of the bell and the thickness of its walls any desired motion can 
be imparted to the bell. The displacement of the bell is a measure 
of the weight of steam flowing and its motion may be transmitted 
through suitable linkage to recording or integrating attachments. 

The "Bailey Boiler Meter" is a combination of the Bailey draft 
gauge (Fig. 577), Bailey steam meter and a recording thermometer. 
By this combination the differences in draft between furnace and ash 
pit, furnace and uptake, temperature of the steam and rate of steam 
flow can be simultaneously recorded on a single clock-driven chart. 
This instrument is compact and easily applied. When correctly in- 
terpreted the records are of great assistance in regulating the rate of 
air supply to the furnace and in controlling the thickness of fire. 

409. Pressure Gauges. — The Bourdon type of gauge, either auto- 
graphic or indicating (Fig. 574), is the most familiar and satisfactory 
means^ of measuring pressures up to 1500 
pounds per square inch or more, although 
diaphragm gauges are also used and both 
are employed as vacuum gauges. For the 
latter purpose, however, the mercurial 
vacuum gauge has the advantage of greater 
accuracy and is not subject to derange- 
ment. Bourdon gauges should be frequently 
standardized by comparison with a gauge 
of known accuracy, a mercury column, or 
a gauge tester. 

For measuring very low pressures, such ^^^- ^*^^' Bourdon Pressure 
as are found in boiler flues or gas mains, ^^^^" 

indicating or recording diaphragm gauges may be had, but some 
form of U-tube manometer is generally employed, the design best 
adapted to the purpose depending upon the accuracy required. The 
simple U-tube (Fig. 575), when filled with mercury, may be used for 
pressures limited only by the inconvenience due to length of tubes, or 
with water as the fluid, for pressures only a fraction of an ounce per 
square inch. Where greater accuracy is required than can be obtained 
with the simple U-tube, some modification may be employed, such as 
the Ellison draft gauge with one incHned leg which magnifies the reading 




826 



STEAM POWER PLANT ENGINEERING 



several times. A form of sensitive gauge is sometimes used which 
depends upon the use of two fluids of different specific gravity, as oil 
and water. 

The Blonck Boiler Efficiency Meter, Fig. 576, consists essentially of 
two differential draft gauges, one connected between the ash pit and 




ORAFT CAUCC 



"^ 



CIMPLC U TUBt 




Fig. 575. Different Forms of Manometer Pressure Gauges. 

furnace and the other between the furnace and the breeching on the 
boiler side of the damper. In the indicating device the lower gauge 
(showing the pressure drop through the fuel bed) is suppfied with red 
colored oil, and the upper gauge (showing the pressure drop between 
furnace and damper) is suppfied with blue colored oil. The readings of 
each gauge and the difference in readings between both gauges are 
indications of the furnace performance and offer a means of scientifi- 
cally controlling the depth of fire, air supply, and rate of combustion. 



A 



JS^ 



^ ^ 



^ 



I0- 




r^T;rpA 





■Id ft' 

BOILER 
EFFIGIENCY 

METER 

T YPE C 
No.l_4a_l 



A.BLONCK4C0 

CHICAGO. ILL. 

si 



^S7 ^27 

Fig. 576. Blonck Efficiency Meter. 

Sliding pointers enable the fireman to fix the draft indications best 
suited for the particular equipment and conditions of operation. This 
device is also made recording. 

A simple yet accurate instrument for measuring and recording very 
low pressures or a small pressure difference is shown diagrammatically 



TESTING AND MEASURING APPARATUS 



827 



in Fig. 577. Two bells A and B are suspended from opposite ends 
of a beam (which is pivoted on knife edge bearings) and are partly sub- 
merged in a light non-volatile oil as indicated. In measuring pressures 
less than atmospheric, connection is made at P2 and Pi is left open to 
the atmosphere. For pressures above atmospheric, connection is made 
at Pi and P2 is left open. For measuring the difference of two pressures 
the higher pressure is applied at Pi and the lower at Pi. If a slight 
suction pressure is applied at P2 it is effective over the inside area of 
bell A and pulls it down into the Hquid. The relative motion between 

bells A and B is transmitted through ^ 

levers L and pen arm P to the re- 
corder pen. This instrument may be 
designed to record pressures or pres- 
sure difference as low as one one- 
thousandth of an inch of water. 

410. Measurement of Temperature. 
— For power-plant purposes mercu- 
rial thermometers are most conven- 
ient for measuring temperatures up 
to 400 deg. fahr., and are inexpen- 
sive. For higher temperature, up to 
say 800 deg. fahr., they are also 
adapted, but must be made of special 
glass and the space above the mer- 
cury filled with nitrogen under pres- 
sure to prevent vaporization of the 
mercury. Such thermometers must 
be used intelligently and should be 
standardized from time to time, since 

they are subject to considerable change. The Bureau of Standards at 
Washington, D. C, is prepared to furnish certificates for which a 
nominal charge is made. 

Fig. 578 shows a form of thermometer which is much used where a 
continuous autographic record is required. It depends for its operation 
upon the pressure produced by a fluid, liquid or gaseous, contained in 
a small bulb and exposed to the temperature to be measured. The 
pressure is transmitted to the recording mechanism through a flexible 
capillary tube which may be of considerable length. Such thermom- 
eters are suital)le for feed water, flue gas, and temperatures not exceed- 
ing 1000 deg. fahr. 

Fig. 579 illustrates a form of electrical pyrometer employing thermo- 
couples which has come into wide use as a reliable means of measuring 




Fig. 577. 



Bailey Recording Draft 
Gauge. 



828 



STEAM POWER PLANT ENGINEERING 




Fig. 578. Bristol Recording Pyrometer. 




Fig. 579. Bristol Thermo-electric Pyrometer. 



TESTING AND MEASURING APPARATUS 



829 



temperatures up to 2600 deg. fahr. The couples most frequently used 
are composed of platinum and platinum-rhodium, platinum and plati- 
num-iridium, copper and copper-constantan, and copper and nickel, 



C 




4 PLATINUM WIRE ICAOS 



w 



Fig. 580. Element for Callendar Resistance Pyrometer. 

the first named being adapted to the higher ranges of temperature. 
The electromotive force set up, when the thermo-j unction is heated, is 
proportional to the temperature and is measured by means of a sensitive 
miUivoltmeter which is usually graduated to read temperature directly. 
Thermo-couples may be made to give an autographic record by means of 
a thread recorder. 

Fig. 580 shows the element of an electrical thermometer based upon 
the change in resistance of a platinum wire when subjected to change 



DIFFUSING GLASS 



*^ FLAME 
QAUQE 



^AMYL-ACETAT 
LAMP 




Fig. 581. Wanner Optical Pyrometer in Position for Standardizing. 

in temperature. The resistance, in terms of temperature, is measured 
by a Whipple indicator, a convenient and portable form of Wheat- 
stone bridge, or may be autographically recorded by means of a Callen- 
dar recorder. Resistance thermometers of this type are very sensitive 



830 



STEAM POWER PLANT ENGINEERING 



and accurate, not easily deranged, and are limited in range only by the 
fusing points of the platinum and the porcelain protecting sheath. 

For higher temperatures and for obtaining the temperatures of in- 
closed spaces above about 900 deg. fahr., such as boiler furnaces, an- 
nealing ovens, and kilns, various forms of optical and radiation pyrom- 
eters have been devised. In such devices no part of the instrument is 
exposed to the temperature to be measured and hence suffers no injury 
from this cause. Optical pyrometers are based upon the measure- 
ment of the brightness of the hot body by comparison with a standard. 
The Wanner optical pyrometer is shown in Fig. 581. After standard- 
izing by comparison with an amyl-acetate lamp, it is only necessary to 
focus the instrument upon the source of heat to be measured and the 
temperature is read on the graduated scale. 

Radiation pyrometers depend upon the measurement of the heat 
radiated from the hot body. The Fery radiation pyrometer, Fig. 582, 



,To Galvanometf.r 




Rack and 
Pinion 



Fig. 582. Fery Radiation Pyrometer. 

is the best-known instrument of this type. When focused upon the 
source of heat a cone of rays of definite angle is reflected by means of 
the mirror upon a thermo-couple located in its focus. The electromotive 
force set up is measured in terms of the temperature of the source of heat 
by a millivoltmeter. Neither the couple nor any part of the instrument 
is ever subjected to a temperature much above 150 deg. fahr. The 
indications are practically independent of the distance from the source 
of heat, and the range is without limit. 

The Uehling pyrometer depends for its operation upon the flow of gas 
between two apertures, thus: Air is continuously drawn through two 
apertures by a constant suction produced by an aspirator. So long as 
the air has the same temperature in passing through these orifices there 
is no change in the partial vacuum in the chamber between them; if, 
however, the air passing through the first opening has a higher tem- 
perature than that passing through the second, the vacuum in the 



TESTING AND MEASURING APPARATUS 



831 



chamber will increase in proportion to the difference in temperature 
since the volume of air varies directly with the temperature. In the 
application of this principle, the first aperture is located in a nickel 
tube which is exposed to the heat to be measured, while the second ap- 
erture is kept at a uniform lower temperature. This style of pyrom- 
eter is made to indicate and record and the indicating and recording 
mechanism can be placed at a distance from the main instrument. 



TABLE 134. 
TYPES OF THERMOMETERS IN GENERAL USE. 









Range in Deg.Fahr. 


Principle of Operation. 


Type. 


for which they 








can be used. 


Expansioa 


.Tiiose depending on the 


Gas 


— 400 to +2900 
-35 to +950 




change in volume or 


Mercurj-, Jena glass, 




length of a body with 


and nitrogen. 






temperature. 


Glass and petrol ether. 


-325 to +100 






Unequal expansion of 


to 950 






metal rods. 




Transpiration and vis- 
cosity. 


Those depending on the 
flow of gases through 


The Uehling 


to 2900 








capillary tubes or small 








apertures. 






Thermo-electric 


.Those depending on the 
electro-motive force 
developed by the dif- 
ference in temperature 
of two similar thermo- 
electric junctions op- 
posed to one another. 


Galvanometric 


— 400 to +2900 


Electric resistance 


.Those utilizing the in- 


Direct reading on indi- 


-400 to+ 2200 




crease in electric resist- 


cator or bridge and 






ance of a wire with 


galvanometer. 






temperature. 






Radiation 


.Those depending on the 


Thermo-couple in focus 


300 to 4000 




heat radiated by hot 


of mirror. 






bodies. 
.Those utilizing the 
change in the bright- 


Bolometer 


— 400 to Sun 


Optical 


Photometric compari-1 
son. 










ness or in the wave 


Incandescent filament 


1100 to Sun 




length of the light 


in telescope. [ 




emitted by an incan- 


Nicol with quartz plate 






descent body. 


and analyzer. 




Calorimetric 


.Those depending on the 


Platinum ball with 


32 to 3000 




specific heat of a body 


water vessel. 






raised to a high tem- 








perature. 






Phisioa 


Those dpnpnrlinp' nn ihc 


Alloys of various fusi- 
bilities. (Seger cones.) 


32 to 3350 




unequal fusibility of 






various metals or 








earthenware blocks of 








varied composition. 







832 STEAM POWER PLANT ENGINEERING 

Table 134 embodies in outline the principles and temperature ranges 
of the various types of thermometers in use. Temperature ranges 
verified by U. S. Bureau of Standards. 

Modern Methods of Temperature Measurements: Cassier's Mag., June, 1909, p. 
99. High Temperature Measurements: Eng. and Min. Jour., Sept. 2, 1911, p. 447; 
Power, Aug. 2, 1910, p. 1376; Engineering, Feb. 9, 1912, Bui. No. 2, Bureau of 
Standards. 

411. Power Measurements. — Instruments for the measurement of power 
may be divided into two general classes, direct and indirect. The former 
involve the direct measurement of force and linear velocity or torque 
and angular velocity and the latter give the equivalent in other forms 
of energy. Direct power measuring appUances include the various 
speed indicators, transmission and absorption dynamometers, and the 
indirect include ammeters, voltmeters, watt-hour meters, boiler flow 
meters, and the like. In all power measurements the time or speed 
factor is readily determined but the force or torque factor, or equivalent, 
often involves considerable labor and the use of costly and complicated 
apparatus. The various conversion factors for the measurement of 
work, power, and duty are given in Appendix F. 

413. Measurement of Speed. — The following chart gives a classi- 
fication of a number of well-known instruments for determining linear 
and angular velocities. 

Hand i Worm and Wheel. 

Counters I [^ ^ . 

Continuous i Gear Train. 

( Electrical. 

Centrifugal j ^^|g^^/- 

Electrical. 

Resonance | Frahm's. 

O-onograph j ^'^'^SnTf or^" 

The most commonly used device for speed determinations is the hand 
speed counter^ consisting of a worm, worm wheel, and indicating dials. 
The errors to be corrected are principally those due to shpping of the 
point on the shaft, and to the slip of the gears in the counting device 
in putting in and out of operation. In some of the better grade of in- 
struments the gears are engaged or disengaged with the point in con- 
tact with the shaft. In the latter design a stop watch, actuated by the 
disengagement gear, minimizes the error likely to occur in hand manipu- 
lation. 

The continuous counter consists of a series of gears arranged to oper- 
ate a set of indicating dials. It may be operated by either rotary or 



Tachometer or 
Speed Indicators . 



TESTING AND MEASURING APPARATUS 833 

reciprocating motion. The rate of rotation is calculated from the read- 
ings of the counter. 

All tachometers indicate directly the speed of the machine to which 
they are attached and are independent of time determination. The 
most commonly used devices depend upon the centrifugal force of re- 
volving weights for their operation. The indicating needle is at- 
tached to the weights in such a manner that the number of revolutions 
per minute is read directly from the position of the needle on the dial. 
These instruments should be caUbrated for accurate work because of 
the number of wearing parts. 

Liquid tachometers consist essentially of small centrifugal pumps dis- 
charging into a vertical tube. The height of the indicating column is 
a function of the speed of rotation. 

Electrical tachometers are miniature dynamos, the voltage being a 
measure of the speed of rotation. These instruments are accurate and 
readily attached but necessitate the use of a delicate and costly volt- 
meter. The indicating mechanism may be placed at any distance from 
the small dynamo and in this respect has a marked advantage over the 
other types of speed incUcators. 

The resonance tachometer affords a convenient method of measuring 
speeds over a wide range. It consists of a number of steel reeds of dif- 
ferent periodicity mounted side by side on a suitable frame. When 
used to measure the speed of an engine or turbine the instrument is 
placed on or near the bed plates and the slight under or over balance 
causes the proper reed to vibrate in unison. 

413. Steam-engine Indicators. — This subject has been extensively 
treated by various authorities and a general discussion would be with- 
out purpose. For indicated horsepower, testing indicator springs, and 
analysis or indicator diagrams see ''Rules for Conducting Steam Engine 
Tests," A.S.M.E. Code of 1915. 

414. Dynamometers. — Dynamometers for measuring power are of 
two distinct types, absorption and transmission. In the former the 
power is absorbed or converted into energy of another form while in 
the latter the power is transmitted through the apparatus without loss, 
except for minor friction losses in the mechanism itself. 

The ordinary Prony Brake is the most common form of absorption 
dynamometer. In the various forms of Prony brakes the power is 
absorbed by a friction brake applied to the rim of a pulley. For low 
rubbing speeds and comparatively small powers it affords a simple and 
inexpensive means of measuring the actual output. 

The Alden absorption dynamometer is a successful form of friction 
brake and has a wide field of application. It has been constructed in 



834 STEAM POWER PLANT ENGINEERING 

large sizes and is adapted to all practical ranges of speed. For a descrip- 
tion of rope brakes and the Alden absorption dynamometers see Ap- 
pendix No. 19, p. 179, A.S.M.E. Code of 1915. 

Water brakes are finding much favor with engineers for high-speed 
service. There are two types, the Westinghouse and the Stumpf. In 
the former the rotor consists of a simple drum with serrated periphery 
revolving in a simple casing, the inner surface of which is serrated in 
a manner similar to the rotor. The resistance is produced by friction 
and impact, and the power is converted into heat which is carried away 
by the circulating water. The casing is free to turn about the shaft 
but is held against rotation by a lever arm. The torque of the lever 
arm is determined as in a Prony brake. A brake of this design, 2 feet 
in diameter and 10 inches wide, will absorb about 3000 horsepower at 
3500 r.p.m. In the Stumpf type the rotor consists of a number of 
smooth disks mounted side by side on a common shaft. The casing is 
divided into a number of compartments corresponding to the division 
of the rotor. There is no contact between rotor and casing. The 
friction between the disks and water and the water and casing tends 
to rotate the latter and the torque is measured in the usual way. In 
either type the power output is readily controlled by the water supply. 

Pump brakes and fan brakes are also used as absorption dynamometers. 
The latter are commonly used in connection with automobile engine 
testing. 

Electromagnetic brakes are occasionally used for power measurements. 
They consist essentially of a metal disk or wheel revolving in a mag- 
netic field. The resistance or drag tends to revolve the field casing 
and the torque is measured in the usual way. 

An electric generator mounted on knife edges forms the basis of the 
Sprague electric dynamometer. The prime mover drives the armature 
of the generator and the reaction between armature and field is counter- 
balanced by suitable weights. The output is conveniently regulated 
by a water rheostat. 

Transmission dynamometers are seldom used for testing prime movers 
and are ordinarily limited to small power measurements. In some in- 
stances, however, as in marine service, transmission dynamometers afford 
the only practical means of approximating the net power deUvered to 
the propeller. For comparatively small power measurements may be 
mentioned the Morin, Kennerson, Durand, Lewis, Webber, and Emerson 
transmission dynamometers, and for large powers, the Denny and 
Johnson electrical torsion meter and the Hopkinson optical torsion 
meter. For detailed descriptions of these appUances consult "Experi- 
mental Engineering, " Carpenter and Diederichs, Chap. X. 



TESTING AND MEASURING APPARATUS 



835 



415. Flue Gas Analysis. — It has been shown (paragraph 22) that 
the products of combustion, commonly called flue gases, resulting from 
the complete oxidation of coal with theoretical air supply consist chiefly 
of nitrogen and carbon dioxide, with lesser amounts of water vapor and 
sulphur dioxide. It was also shown that with a deficient air supply 
the flue gases may contain carbon monoxide and varying amounts of 
hydrocarbon. If excess air was used in the combustion of the fuel free 
oxygen would be present in the gases. Evidently an analysis of the 
flue gases offers a basis for judging the efficiency of combustion. The 
first step in the analysis and the most important one is the obtaining 
of a representative sample. Since the gases in the breeching and flues 
may be far from homogeneous great 
care must be exercised in getting a 
true average sample. (See Appara- 
tus and Methods for Sampling and 
Analysis of Furnace Gases, U. S. Bu- 
reau of Mines, Bui. No. 12, 1911.) 

The analysis as ordinarily made 
in commercial practice is called vol- 
umetric, although in reality it is 
based upon the determination of 
partial pressures. According to 
Dalton's laws when a number of 
gases are confined in a given space 
each gas occupies the total volume 
at its own partial pressure, and the 
total pressure is the sum of all 
the partial pressures. When one of 
the gases is absorbed by a suitable 
medium and the remaining gases 
are compressed back to the original total pressure, a volume decrease is 
found, and if the temperature remains constant this decrease represents 
the volume absorbed. 

The apparatus usually employed for volumetric analysis consists of 
a graduated measuring tube into which the gases are drawn and accu- 
rately measured under a given pressure, and a series of treating tubes, 
containing the necessary absorbing reagents, into which they are trans- 
ferred until absorption is complete. The Orsat apparatus, Fig. 583, 
forms the basis of nearly all of the portable appliances on the market 
for analyzing flue gases and the ordinary products of combustion. In 
this apparatus a measured volume, representing an average sample of 
the gas, is forced successively through pipettes containing solutions of 




Fig. 583. Standard Orsat Apparatus 
for Flue Gas Analysis. 



836 



STEAM POWER PLANT ENGINEERING 



caustic potash, pyrogallic acid and cuprous chloride in hydrochloric 
acid, respectively, thus absorbing the carbon dioxide, the oxygen and 
the carbon monoxide, the contraction of volume being measured in each 
case. The apparatus as originally constructed is 
bulky and fragile and slow in its absorption of gas. 
The Hempel Apparatus works on the same prin- 
ciple as the simple form of Orsat apparatus de- 
scribed, so far as the latter is applicable, excepting 
that the absorption may be hastened by shaking 
the pipettes bodily, bringing the chemical into 
most intimate contact with the gas. It is less port- 
able and in some particulars it requires more care- 
ful manipulation than the Orsat, while for general 
analysis it is not adapted unless used in a well- 
equipped chemical laboratory. The absorption pipettes are made in 
sets which are shaped in the form of globes, and a number of independ- 
ent sets are required for the treatment of the different constituent gases. 
A simple pipette of the Hempel type is shown in Fig. 584. 

The Williams Improved Gas Apparatus is a marked improvement 
over the standard Orsat in that the objections cited above are obvi- 
ated. In addition to the elimination of these objectionable features 





Fig. 585, Williams Improved Gas Apparatus. 



provision is made in the ''Model A" type for the determination of 
illuminants, hydrogen and methane along with the three gases mentioned 
above. Referring to Fig. 585, A, B, C, and D are pipettes containing 



TESTING AND MEASURING APPARATUS 



837 



the necessary reagents for absorl)ing, respectively, CO2, illuminants, O2, 
and CO. M is a graduated measuring flask connected at the bottom 
with water-level })ottle W and at the top with the various pipettes. F 
is a hard rubber pump for taking gas sample directly from the source of 
supply, thereby eliminating the inaccuracy and annoyance of collection 
over water and transference. P is a portable case containing a spark 
coil and batteries for exploding the methane and hydrogen remaining 
in the burette after the other constituents have been removed. When 
extreme accuracy is desired mercury is used as the displacement medium 
in the leveling bottle since water absorbs CO2 to a certain extent. For 
a complete description of this apparatus with sample calculations see 
paper read by F. M. Williams before Division of Industrial Chemists 
and Chemical Engineers, American Chemical Society, Washington, 
D. C, Dec. 28, 1911. 

"Little" Modified Orsat Apparatus. — Fig. 586 illustrates a modified 
Orsat apparatus as used by the Arthur D. Little Company of Boston, 




III iiiii ; 

illlllr 



Fig. 586. Modified Orsat Apparatus. — Arthur D. Little Co. 

Mass. The right half of the device is the ordinary Orsat apparatus 
and the left portion constitutes the sampHng attachment. The gases 
are drawn from the source of supply through rubber tube (2) into the 
sampling pipette (3) and out through rubber tube (1) to the aspirator. 
The latter may be operated by steam or water. When a sample is 



838 



STEAM POWER PLANT ENGINEERING 



being collected the three-way cock on the glass header is closed and the 
mercury in the sampling tube (4) is allowed to drain through the movable 
overflow into the mercury retainer. The overflow is lowered at a 
constant rate by clockwork. Two driving pulleys afford seven different 
rates of movement downward of the overflow, thereby enabling a con- 
tinuous sample to be collected at constant rate over any period from 



to 24 hours. 



Instantaneous samples may be drawn off and analyzed 
as often as desired and with 
practically no delay to the con- 
tinuous sample. For further 
details see Power, Julv 16, 1912, 
p. 77. 

For many practical purposes 
it is sufficient to determine the 
carbon dioxide. A number of 
satisfactory appliances are on 
the market which give continu- 
ous autographic records of the 
percentage of CO2 on clock- 
driven charts. These devices, 
however, are rather expensive 
and usually beyond the appro- 
priation of small boiler plants. 

Simmance-Abady CO2 Re- 
corder. — Fig. 587 illustrates the 
general principles of the Sim- 
mance-Ahady CO2 Recorder. The 
operation is as follows: A con- 
tinuous stream of water enters 
reservoir K through inlet X 
and overflow at 0. A portion 
of the stream flows into tank 
A through pipe F and 




Simmance-Abady CO2 Recorder. 



Fig. 587. 

" ' causes 

bell float B to rise. As the float rises it permits bell D of the ex- 
tractor to fall. When float B reaches the top of its stroke it raises 
valve stem E, trips the valve and causes the water to siphon out 
of tank A through siphon tube G. The lowering of the water level 
allows the bell to sink. As it falls it draws up the water-sealed 
extractor bell D and creates a partial vacuum under the latter. Flue 
gas then flows from the source of supply through P and H into the bell. 
The mass of water discharged from siphon tube G into the small vessel 
V beneath it overcomes the counterweight Q and closes the balance 



TESTING AND MEASURING APPARATUS 



839 



To Boiler Room Indioator 
To Recording Gauge 



Caa8ticJ>rip 



Absorption Chamber 



valve H, thereby entrapping a fixed volume of gas in the extractor bell. 
The stream of water which is continually flowing into tank A causes the 
float B to rise and the bell D to sink, as before. The lowering of bell D 
forces the entrapped flue gas through the caustic potash solution in 
vessel M into water-sealed recorder bell /. The displacement of bell 
J will be less than that of beU D by the volume of CO2 absorbed in 
vessel M. The percentage of CO2 in the flue gas is thus indicated by 
the position of the bell J with reference to the graduated scale A^. The 
pen mechanism is attached to bell J and records the percentage of C()2 
by the length of lines on a clock-driven chart. These samples are 
analyzed and the lines are drawn at three-minute intervals. The small 
water aspirator at X is for the pur- 
pose of exhausting gas continuously 
from the pipes connecting the re- 
corder to the boiler, thereby insuring 
true samples at the time of absorp- 
tion. Auxiliary pipe P is connected 
to main gas lead P. 

The Uehling Composimeter is an- 
other successful instrument for con- 
tinuously recording the percentage of 
CO2 in the flue gas. The principles 
of this apparatus are illustrated in 
Fig. 588. The device consists pri- 
marily of a filter, absorption cham- 
ber, two orifices, A and B, and a 
small steam aspirator. Gas is drawn 
from the usual source by means of 
the aspirator through a preliminary 
filter located at the boiler, and then 
through a second filter as illustrated 
in the diagram. From the latter the gas passes through orifice A, 
thence through the absorption chamber and orifice B to the aspirator 
where it is discharged. The CO2 is absorbed by the caustic potash 
solution in the absorption chamber. This reduces the volume and 
causes a change in tension between the two orifices in proportion to 
the CO2 content of the gas. This variation in tension is indicated by 
the water column, as shown, and is transmitted by suitable piping to 
the recording mechanism which may be placed at a considerable dis- 
tance from the boiler room. 

416. Moisture in Steam. — Several forms of calorimeters are avail- 
able for determining the quafity of steam. The simplest as well as 




-Eiltet 



Oae Inleb 



Fig. 



ii Caustic Overflow 



588. Principles of the Uehling 
Gas Composimeter. 



840 



STEAM POWER PLANT ENGINEERING 



the most satisfactory, if the percentage of entrained moisture is not 
beyond its range, is the throttling calorimeter, Fig. 589. In this device 
the sample of steam, which is taken from the steam pipe by means of 
the perforated nipple, is allowed to expand through a very small orifice 
into a chamber open to the atmosphere. The excess of heat hberated 
serves first to evaporate any moisture present and then to superheat 
the steam at the lower pressure. From the observed temperature and 
pressures it is easy to calculate, with the aid of steam tables, the per- 
centage of moisture in the original sample. 

The limit of the throttle calorimeter depends upon the steam pressure 
and is about 3 per cent of moisture at 80 pounds pressure and about 



-Thermomete]: 




To Atmosphere- 

Fig. 589. A Typical Throttling Calorimeter. 

5 per cent at 200 pounds. For steam containing greater percentages 
of moisture the separating calorimeter, Fig. 590, is sometimes used. 
This instrument is virtually a steam separator and mechanically sepa- 
rates the moisture from the sample of steam. The water thus separated 
collects in a reservoir provided with gauge glass and graduated scale, 
while the dry steam passes through an orifice to the atmosphere. The 
weight of dry steam per unit of time is indicated on the gauge, calcu- 
lated according to Napier's rule, or may be determined by condensing 
and weighing. The accuracy of the moisture determination is greatly 
affected by the difficulty of obtaining true samples of steam containing 
large percentages of moisture. 

Fig. 591 shows the Ellison universal steam calorimeter, which com- 
bines the separating and throttling principles and is adapted to steam 
of any degree of wetness. The separating chamber is provided with 



TESTING AND MEASURING APPARATUS 



841 



a gauge glass, not shown, for indicating the weight of water which 
accumulates only when the steam is too wet to be superheated. 

Throttling Calorimeters: Power, Dec, 1907, p. 891; Trans. A.S.M.E., 17-151; 
175, 16-448; Engr. U. S., Feb. 15, 1907, p. 219. 

Separating Calorimeters: Trans. A.S.M.E., 17-608; Engr. U.S., Feb. 15, 1907, 
p. 219. 

Universal Calorimeter: Trans. A.S.M.E., 11-790. 

Thomas Electrical Calorimeter: Power, Nov., 1907, p. 791. 




DRAIN COCII 



Fig. 590. Carpenter Separating 
Calorimeter. 



Fig. 591. Ellison Universal Steam 
Calorimeter. 



417. Fuel Calorimeters. — The analysis and heat evaluation of 
fuel require considerable time and skill and much costly apparatus, 
hence in most power plants it is customary to depend upon a specialist 
to whom samples are submitted from time to time. In many large 
stations, however, the conditions often warrant the establishment of a 
testing laboratory equipped for the proximate analysis of coal and the 
determination of the calorific value of the solid, liquid or gaseous fuel 
used. The Mahler bomb calorimeter illustrated in Fig. 592 is the 
most accurate and satisfactory device for solid and liquid fuels but is 
comparatively expensive. The instrument consists of a steel shell or 



842 



STEAM POWER PLANT ENGINEERING 



"bomb" of great strength, lined with porcelain or platinum, into 
which a weighed sample of the fuel is introduced and burned on a 
platinum pan in the presence of oxygen under a pressure of about 300 



A Insulation 

B Bomb 

C Platinum Pan 

D Water 

E Electrode 

F Ignition Wire 

G Stirring Device 

S Support for Stirrer 

T Sensitive Thermometer 

O Oxygen Tank 




Fig. 592. Mahler Bomb Calorimeter. 

pounds per square inch. The charge is ignited by an electric current. 
During combustion the bomb is submerged in a known weight of water 
which is kept constantly agitated. The calorific value is calculated 

from the observed rise in tempera- 
ture due to the heat evolved, proper 
corrections being made for the water 
equivalent of bomb and appurte- 
nances, heat given up by the ignit- 
ing current, and for radiation or 
absorption of heat from the sur- 
rounding air. 

The Parr calorimeter. Fig. 593, 
is an inexpensive instrument, very 
simple in operation, and gives re- 
sults which are sufficiently accurate 
for all practical purposes. The 
weighed sample of coal, together 
with a quantity of sodium peroxide 
which supplies the oxygen for com- 
bustion, is introduced into the car- 
tridge. Means are provided for rotating the cartridge when submerged 
in the calorimeter, the attached vanes agitating the water to maintain 
uniform temperature. The charge is fired either electrically or by 




Parr Fuel Calorimeters. 



TESTING AND MEASURING APPARATUS 



848 




Fig. 594. Individual Boiler 
Control Board. 



introducing a short piece of hot wire through the conical valve. The 
calorific value is calculated from the o]:)served rise in temperature and 
the constants of the instrument. Among other forms of instruments, 
in more or less general use and which give very satisfactory results, 
may be mentioned the Carpenter, Thomp- 
son, Atwater and Emerson calorimeters. 

Comparison of Different Types of Calorimeters: 
Jour. Soc. Chem. Ind. (1903), 22-1230 

418. Boiler Control Boards. — In the mod- 
ern large central station efficient operation 
of the various units composing the plant is 
greatly facihtated by grouping the testing 
instruments on a control board and by 
placing this board where it can be conven- 
iently studied by the operating engineer. 
Fig. 594 shows the individual control board 
as installed before each boiler unit in the 
Northwest plant of the Commonwealth Edison Company of Chicago, 
and Fig. 595 shows the section control board for each turbine unit. 
The individual control board is mounted on the front of the boiler 
casing and the section board is placed at the end of the battery of 

boilers near the wall dividing the boiler 
from the turbine room. With reference to 
Fig. 594 the two instruments at the top are 
steam flow meters — one on each steam lead 
— with indicating, recording and integrat- 
ing attachments. These meters show the 
amount of steam dehvered at any time by 
the boiler and gives a complete record of its 
delivery. The three recording gauges be- 
low show the temperature in uptake from 
the boiler, the temperature of the feed 
water leaving the economizer and entering 
the boiler and the temperature of the flue 
gases leaving the economizers. Below and at 
the left is a CO2 recorder, while at the right- 
hand corner are two indicating draft gauges, 
one connected to the furnace and the other to the uptake. With 
reference to the section control board, the two flow meters at the top 
measure the steam input to the turbine and the feed water input to 
the boilers, respectively. The recording thermometers immediately 



00 

o 

000 



Fig. 595. Boiler Section 
Control Board. 



844 STEAM POWER PLANT ENGINEERING 

below show the temperature of the steam entering the turbine and the 
temperature of the feed water entering the economizer, respectively. 
Below these are two recording pressure gauges showing the pressure on 
the steam header and on the boiler feed header, respectively, while in 
the center of the board is a clock and below that an indicating watt- 
meter showing the output of the turbo-generator unit which is direct 
connected to these boilers. Where automatic coal-weighing devices are 
in use the individual control board includes the fuel measuring dials. 
By the use of these instruments a very complete check is obtained of 
the performance of individual boilers and of the entire unit. 



CHAPTER XVIII 

FINANCE AND ECONOMICS. — COST OF POWER 

419. General Records. — In all power plants, public or private, an 
itemized record of plant performance and cost of operation is of vital 
importance for the most economic results. In many states public 
utility corporations are required to submit an annual statement cover- 
ing the various details of operation, and in order to insure uniformity 
ruled and printed forms are furnished by the state. The private plant 
owner, on the other hand, is free to use his own judgment and may 
adopt any system of cost accounting or dispense with them entirely. 

The principal objects of keeping a system of records are (1) to enable 
the owner to compare the performance of his plant from time to time 
and to show him exactly what his plant is costing him, and (2) to enable 
the engineer to analyze the various records with a view of reducing all 
losses to a minimum. Power-plant records to be of value must be 
closely studied with a view to improvements. The mere accumula- 
tion of data to be filed away and never again referred to is a waste of 
time and money. 

Records should cover not only the daily, monthly, and yearly oper- 
tion of the plant but also, as permanent statistics, a complete analysis 
of each item of equipment. The value of such data cannot be over- 
estimated. The engineer will frequently find it greatly to his interest 
to have available at a moment's notice the complete details of his 
engines, boilers, generators, and other machinery, especially when it is 
required to renew a broken or worn-out part in case of emergency. 

The question of whether to purchase power or to generate it depends, 
chiefly, upon the relative cost of the two methods, although the absence 
of power-plant machinery and freedom from the coal- and ash-handling 
nuisance may be important factors. There is no doubt that the cen- 
tral station can generate power cheaper than the small isolated plant, 
but in most cases it is a question not only of power, but also of 
supplying steam for heating and other purposes, and a careful study of 
all of the items entering into the problem is necessary for an intelligent 
choice. The service department of the large central station with its 
carefully maintained system of records has a strong advantage in pre- 
senting its arguments over the average private plant with its ill-kept 
and faulty system of accounting, and in some instances central-station 

845 



846 



STEAM POWER PLANT ENGINEERING 



service has been adopted simply because the engineer in charge was not 
in a position to prove positively that his own plant was the better 
investment. R. J. S. Pigott, Jour. A.S.M.E., Dec. 1916, p. 947, shows 
the effects of modifying the operating conditions of power plants, and 
of changing the character of the auxiliary equipment by means of 
graphic analysis. From the study of such an analysis the cost of 
producing power for given conditions may be determined with little 
effort, and the effects of changes in the conditions or equipment may 
be predetermined with accuracy. 



TABLE 135. 
PERMANENT STATISTICS. 

General Information. 



Date of installation 

Type of building 

Number of floors 

Number of offices 

Volume of building, cu. 

ft 

Type of heating system . 

Engine room, sq. ft 

Height of chimney, ft. . . 
Draft, inches of water . . 

Kind of grate or stoker . 

Kind of coal Ill 

Coal storage capacity, 
tons 

Capacity ice plant, tons 
in 24 hrs 

Capacity storage bat- 
tery, am. hrs 



Office 

18 
900 

10,860,000 

Webster 

6,840 

318 

3.5 

Jones 

Underfeed 

screenings 

450 

50 

None 



Total cost of building. . . 

Ground plan 

Total office floor space, 

sq. ft 

Height of building 

No. of sides exposed. . . . 
Radiator surface, sq. ft . 

Boiler room, sq. ft 

Number of elevators. . . . 

Type of elevators j 

Capacity of elevators, 
lb,, each 

Boiler pressure 

Back pressure 

Part of bldg. lighted. . . . 

Total cost of mechanical 
plant 



$5,000,000 
191X231 

400,000 

280 

3 

100,000 

5,400 

22 

High pressure 

hydraulic 

2,700 

150 

Atmospheric 

All 

$650,000 





Engines. 


Generators. 


Motors. 


Boilers. 


Type.... 

Number installed 


Ball compd. 

5 

250 h.p. 


Crocker 
5 


-Wheeler 
25 


5 


Rated capacity 


150 kw. 




375 h.p. 





LIGHTS. 








Incandescent. 


Arcs. 


Type 


Carbon 
150 


Tungsten 
30,000 


Inclosed 


Number installed 


15 



A number of attempts have been made to standardize power-plant 
records but the results have been far from satisfactory because of the 
wide range in operating conditions. Each installation is a problem 
in itself and the items to be recorded must necessarily depend upon the 



FINANCE AND ECONOMICS — COST OF POWER 847 

size and character of the plant. A common mistake is to attempt too 
comprehensive a system with the result that after the novelty has 
ceased the labor of making the various entries becomes irksome, many 
of the items are omitted, guesses are substituted in place of actual 
observations, and the records are ultimately without value. A few 
properly selected items, accurately recorded, are of vastly more impor- 
tance than an elaborate system of records indifferently maintained. 

Walter N. Polakov, Jour. A.S.M.E., Dec, 1916, p. 966, has proposed 
a ''standardization of power plant operating cost" by means of which 
the owners of power plants can judge, without the necessity of going 
into technical details themselves, how close the actual performance of 
the plant is to the possible minimum cost at any time or under any 
circumstances, all variable factors beyond operating control being 
automatically adjusted. Mr. Polakov shows the futility of attempting 
to judge any one plant by the performance of others having a different 
kind of equipment or of a different nature of service. Even where 
conditions appear identical such comparisons do not offer a true meas- 
ure of excellence. It is not so important to know that one's plant is 
better than another as to know whether it is as good as it can be. 
Mr. Polakov shows how this can be determined by the use of curves 
of ''standard costs" the plotting and application of which are explained 
in his paper before the American Society of Mechanical Engineers. 

420. Permanent Statistics. — Tables 135 to 138 are taken from the 
records of a large isolated station in Chicago and serve to illustrate the 
make-up of the "permanent statistics." The complete file covers each 
item of equipment and includes the various drawings, specifications, 
and guarantees for the entire mechanical equipment. Since these 
records do not vary with the operation of the plant they require no 
further attention, once they are compiled, except of course for such 
changes as may be made from time to time in the plant itself. 

420a. Operating Records. — The operating records of any plant bear 
the same relationship to the economical operation of that plant as the 
bookkeeping and cost accounting system bears to the manufacturing 
plant. The distribution of profit and loss in either case can only be 
obtained by itemizing the various factors involved and by grouping 
them in such a manner as to show at any time where improvement is 
possible. Commercial bookkeeping has been more or less standardized 
and entails very little need of originality on the part of the bookkeeper, 
but the selection and maintenance of a system of power-plant records 
may require considerable study and experimenting, since each installa- 
tion is a problem in itself. The items included in the different forms 
depend upon the apparatus provided for weighing the coal and water, 



848 



STEAM POWER PLANT ENGINEERING 



TABLE 136. 



PERMANENT STATISTICS. 



Boilers. 



Make of boiler Stirling 

Total number in plant 5 

Date of installation 

Steam pressure, gauge 

Safety-valve pressure 

Type of safety valve 

Area of grate, sq. ft 

Heating surface, sq. ft 

Superheating surface, sq. ft 
Number of steam drums . . . 
Diameterof steam drums, in. 
Distance between steam 

drums, ft 

Thickness of shell, in 

Thickness of head, in 

Diameter of steam nozzle, 

in 

Diameter of safety valve. . . 
Diameter of blow-off, in — 
Diameter of feed pipe, in. . . 
Temperature of flue, deg. 

Fah 

Temperature of feed water, 

deg. Fah. 

Ratio of heating surface to 

grate area 

Kind of fuel 

Carterville, 111., Screenings 

Type of grate Green chain grate 

Rated horse power 375 



150 

160 

Pop 

3,500 

None 

3 

36 



10 

2-4 in. 

2.5 

2 

450-490 

210 

41.6 



Number in battery 1 

Weight of boiler 62,186 

Cost of boiler and fittings 

(each) $5,400 

Height of setting 17 ft. 9 in. 

Length of setting 17 ft. 4 in. 

Width of setting 15 ft. 3 in. 

Weight of setting 272,000 

Thickness of wall 

Side 20 in.; back, 15 in. 

No. of bricks, fire 6,590 

No. of bricks, common 19,600 

Dimensions of foundation 

15 ft. 2 in. X 17 ft. 4 in. 
Material of foundation 

Stone and concrete 
Cost of foundation and set- 
ting (each) $1,500 

Distance between batteries 4 ft. 6 in. 

Distance back of boiler. ... 17 ft. 6 in. 

Distance in front of boiler. . 16 ft. 6 in. 

Distance overhead 2 ft. 10 in. 

Number of tubes 337 

Diameter of tubes, in 3.25 

Length of tubes, ft 12 to 14 

Steam space, cu. ft 96 

Water space, cu. ft 643 

Kind of draft Forced 

Inches of draft (maximum) . 3.5 



TABLE 137. 

PERMANENT STATISTICS. 

Feed Pumps. 



Date of installation. 

Make 

Number in plant. . . . 

Height, ft 

Length, ft 

Width, ft. 



Snow 

2 

3 

12 

4 

Weight of pump 5 tons 

Cost, each $965 

Steam pressure 150 

Back pressure ^ 

Number of valves 32 

Character of valves 

Rubber, brass lined 
Area thro' valve seats, sq. in., 

per pump 12. 13 

Gallons of water per min., per 

pump 800 

Pounds of water per 24 hrs., 

average, actual 479,400 

Gallons of water per 24 hrs. . . 599.2 

Volume of air chamber, cu. ft. 3 

Shop number 24,572-3 



Diameter of steam cylinder. . 16 

Diameter of water cylinder . , 10 

Stroke 12 

Displacement per stroke, cu. 

ft 0.5454 

No. of strokes per min., aver- 
age 12 

Diameter of suction 8 

Diameter of discharge 5 

Diameter of steam pipe 2.5 

Diameter of exhaust 4 

Diameter of steam drips ^ 

Diameter of water drains. .. . 5 

Suction head, lb. per sq. in.. . If 

Discharge head, lb. per sq. in. 175 
Kind of piston packing 

Outside packed plunger 

Size of piston packing 

Kind of rod packing Soft 

Size of rod packing f 

Temperature of feed water . . . 214 



r 



FINANCE AND ECONOMICS — COST OF POWER 849 

the type and number of instruments available for measuring tempera- 
ture, pressure, and power, and the system adopted for keeping track of 
oil, waste, general supplies, and repairs. In large stations autographic 
recording and integrating appliances, which are to be found in nearly 
all strictly modern stations and represent but a small part of the first 
cost of the plant, greatly reduce the labor of keeping continuous records. 
In small plants the cost of autographic instruments may prove to be 
prohibitive and recourse must be had to the usual indicating devices. 
In the latter case, continuous records may be closely simulated by 
plotting the readings of the indicating appliances, say every 15 minutes, 
or even once ev^ry hour, and by connecting the points with a straight 
line. (See Figs. 601 to 606.) The oftener the readings are taken the 
smaller will be the error. Total quantities may be obtained by sum- 
ming up the various items or by integrating the graphical chart by 
means of a planimeter. It is not sufficient to record monthly or yearly 
averages. Daily and even hourly records are absolutely essential for 
maximum economy. The various losses may be reduced to a minimum 
only by an intelligent analysis of daily records. A number of forms 
taken from the files of various power plants are reproduced in this 
chapter under the proper subheadings and serve to illustrate current 
practice. 

Power Plant Records: Prac. Engr. U. S., Jan. 1, 1914, p. 80; March 1, 1912, p. 242; 
Jan. 1, 1912, p. 36; Power, May 28, 1912, p. 758; Nov. 11, 1913, p. 697. 
Log Sheets at Delray Station: Power, Oct. 5, 1915, p. 182. 

431. Output and Load Factor. — The output of a plant is usually 
stated in terms of the (1) average horsepower, or equivalent, for a given 
period of time. (2) Unit output — horsepower-hours, or equivalent. 

When the plant is operating at practically constant load it is suffi- 
ciently accurate for most purposes to express the output in horsepower, 
or equivalent, per month or per year. When the output fluctuates as 
is the general case, it is best expressed in terms of unit output. For 
example, one horsepower per year, 24 hours per day, and 365 days per 
year is equivalent to 365 X 24 = 8760 horsepower-hours. If the full 
power is used throughout this time it matters little whether the charge 
is based on the flat rate (horsepower per year) or the unit rate (horse- 
power-hours) ; if, however, the power is used only half the time, the 
yearly cost per horsepower-hour will be just double. 

The yearly load factor or simply load factor is the ratio of the actual 
yearly output to the rated yearly output measured on the twenty-four- 
hour basis. Thus : 

J 1 ^ , _ Yearly output, horsepower-hours or equivalent 
Rated horsepower, or equivalent X 8760 



850 STEAM POWER PLANT ENGINEERING 

The curve load factor or station load factor is the ratio of the yearly out- 
put to the rated output based upon the number of hours the plant is in 
actual operation. Thus, for an electric station: 

Yearly output, kilowatt-hours 



Curve load factor 



Rated capacity X hours plant is in operation 



Much confusion arises from the interpretation of the term "rated 
capacity." If rated below the maximum load it can sustain it is evi- 
dent that a prime mover may operate with a load factor over 100 per 
cent, in which case the term is without purpose. The accepted defi- 
nition of rated load in this connection is the maximum load which the 
prime mover can sustain continuously on a twenty-four-hour basis 
without overheating. Other definitions have been assigned to the 
term load factor and station factor, but the two stated above are more 
in accord with current practice. 

In any plant the great desideratum is a high load factor with greatest 
return on the investment. All the factors of expense included in the 
cost of power- are then operating at maximum economy. High peak 
loads and low average loads necessitate large machines which are but 
little used and greatly increase the fixed charges. 

The demand factor is the ratio of the maximum demand to the con- 
nected load. There is a general tendency to overestimate the maxi- 
mum electric demand, due, in a measure, to the possibilities of all the 
lights and motors being in use at one time. Practically speaking, such 
conditions are not likely to occur. Table 139 gives an idea of the value 
of the demand factor for various classes of service and may be used as 
a guide for determining the size of prime movers. 

The diversity factor may be defined as the ratio of the sum of the 
individual maximum demands of a number of loads during a specified 
period to the simultaneous maximum demand of all these same loads 
during the same period. If all the loads in a group impose their 
maximum demands at the same time, then, the diversity factor of 
that group will be unity. See Diversity and Diversity Factors, Terrell 
Croft, Power, Feb. 6, 1917, p. 171. 



FINANCE AND ECONOMICS — COST OF POWER 



851 



TABLE 138. 

LOAD FACTORS — LARGE STATIONS. 



Plant. 



Buffalo General Electric Company 

Cleveland Electric Light Company . . . . 

Duquesne Light Company 

Edison Companies: 

Boston 

Brooklyn 

Commonwealth 

Detroit 

New York 

Southern California 

Minneapolis General Electric Company 

.Philadelphia, Electric Company 

Public Service, N.J 



Peak Load, 

Kilowatts 

(Thousands). 



65.5 

85.0 
101.1 

80.5 

67.2 

369.7 

130.2 

254.8 

60.9 

43.6 

142.3 

174.0 



Yearly Output, 

Kw-hr. 

(Millions). 



299.3 
340.6 
463.5 

238.5 
233.4 
1341.9 
546.9 
856.4 
300.0 
171.6 
444.8 
608.0 



Yearly Load 
Factor, 
Per Cent. 



57.0 

45.8 
52.3 

33.7 
38.1 
43.2 
47.8 
38.3 
56.0 
44.9 
35.6 
39.8 



TABLE 139. 

CENTRAL STATIONS, DEMAND FACTORS. 

Demand factors compiled by Commonwealth Edison Company of Chicago. 

Class of Service. 



Lighting customers: 

Billboards, monuments, and department stores 

Offices 

Residences and barns 

Retail stores 

Wholesale stores 

Average 

Motor customers: 

Offices 

Public gathering places and hotels 

Residences and barns 

Retail stores 

Wholesale stores and shops 

Average 



Demand Factor. 


85.6 


72.4 


60.0 


66.3 


70.1 


59.8 


65.1 


28.7 


69.3 


61.2 


58.2 



59.4 



852 



STEAM POWER PLANT ENGINEERING 



TABLE 140. 

MAXIMUM DEMAND TABLE FOR INSTALLATIONS UNDER ONE KILOWATT 

CONNECTED LOAD. 

Commonwealth Edison Co. 



Connected Load. 


Residence Lighting. 
(Monthly Basis.) 


Commercial Lighting. 
(Monthly Basis.) 


Number of 
Sockets. 


Wattage. 
Equivalent. 


Kw-hr. at 
Full Rate. 


Estimated Maxi- 
mum Number of 

Sockets Used 
Simultaneously. 


Kw-hr. at 
Full Rate. 


Estimated Maxi- 
mum Number of 

Sockets Used 
Simultaneously. 


1 


50 


2 


1 


2 


1 


2 


100 


3 


2 


3 


2 


3 


150 


5 


3 


5 


3 


4 


200 


6 


4 


6 


4 


5 


250 


7 


5 


8 


5 


6 


300 


8 


5 


9 


6 


7 


350 


9 


6 


10 


6.7 


8 


400 


10 


6.7 


11 


7.3 


9 


450 


10 


6.7 


12 


8 


10 


500 


11 


7.3 


13 


8.7 


11 


550 


11 


7.3 


14 


9.3 


12 


600 


12 


8 


15 


10 


13 


650 


12 


8 


16 


10.7 


14 


700 


13 


8.7 


17 


11.3 


15 


750 


13 


8.7 


18 


12 


16 


800 


14 


9.3 


19 


12.7 


17 


850 


14 


9.3 


20 


13.3 


18 


900 


15 


10 


21 


14 


19 


950 


15 


10 


22 


14.7 



In alternating-current motor installations the Wright Maximum 
Demand Indicator is not applicable, so that the maximum demand is 
determined, except in special cases, by the following percentages of 
the rated capacity of the connected load: 

Per Cent. 

Where installations are under 10 hp. and only one motor is 

used 85 

Where installations are under 10 hp. and more than one motor 

is used 75 

Where installations are from 10 hp. to 49 hp., both inclusive 

(irrespective of number of motors) 65 

Where installations are 50 hp. or over (irrespective of number 

of motors) 55 



FINANCE AND ECONOMICS — COST OF POWER 853 



TABLE 141. 

TYPICAL OPERATING CHART. 

DAILY POWER-HOUSE REPORT. 

The United Light and Power Co. 

Division 



Weather — Noon 



Hr. 



Engine No. 1 
Engine No. 2 
Inc. current on 
Street arcs on 



started M 

started M 

M 

M 



stopped M 

stopped M 

off M 

off M 



Total time run . 
Total time run. 
Total time on . 
Total time on . 



Min. 



Noon AMPERE READINGS. 


12 00 


12 30 


1 00 


1 30 


2 00 


2 30 


3 00 


3 30 


4 00 


4 15 


4 30 


4 45 


5 00 


5 15 


5 30 


5 45 


6 00 


6 15 


6 30 


6 45 


7 00 


7 15 


7 30 


7 45 


8 00 


8 15 


8 30 


8 45 


9 00 


9 15 


9 30 


9 45 


10 00 


10 30 


11 00 


11 30 


12 00 


1 00 


2 00 


3 GO 4 00 


5 00 


5 15 


5 30 


5 45 


6 00 


6 15 


6 30 


6 45 


7 00 


7 15 


7 30 


7 45 


8 00 


9 00 


10 00 


11 GO 



Coal used 

Cylinder oil 

Engine oil 

Waste 

Water 

Carbons 

Globes outer. . . .inner 



..lb. 
..pt. 
..pt. 
..lb, 
cu. ft. 



Coal Received on Track. 

Car No 

Initial 

Time placed m 

Time released m 

Weight lb. 



Boilers in Service. 

No. 1 from m to m 

No. 2 from m to m 

No. 3 from mto m 

Washed No 

Blew No 



Ashes sold loads to . 



Material Received for Power House Use. 



Total Kilowatt Output- 
Read meter 12 o'clock noon 



Meter to-day Kw. 

Meter yesterday Kw. 

Diff 



Report here ANY interruption of service either arc or incandescent. 

[ Time off Cause 

Arc lights out 

Lights. . , 



Location Reported by 



854 



STEAM POWER PLANT ENGINEERING 



TABLE 142. 
TYPICAL OPERATING CHART. 
(Large Chicago Department Store.) 
Monthly Report. 



.19. 





Average 
Outside 
Tempera- 
ture. 


Fuel. 


Supplies. 


Date. 


Coal. 


Ash. 


Oil Used, Gals. 


Waste 
Pounds. 


Total 




Kind. 


Pounds 
Burned. 


Cost 
Per 
Ton. 


Cost 
Per 
Day. 


Pounds 
Removed. 


Engine. 


Cylinder. 


Water to 
Building, 
Cu. Ft. 





Output. 






Engine-Hours Run. Boilers-Hours Run. 




Breeching. 


Boilers. 


Generators. 


1 


2 


3 


4 


5 


1 


2 


3 


4 


5 


6 


Draft. 




Pounds 

of 
Water 
Evapo- 
rated. 


Water 
Evapo- 
rated 
Per Lb. 
of Coal. 


Ampere 
Hours. 


Kilo- 
watt- 
Hours. 


Tem- 
pera- 
ture. 



Heating System. 


Ventilating 

Plants, Hours 

Run. 


Refrigerating Plant. 


Repairs- Hours. 


Steam 
Pressure. 


Live 

Steam- 
Hours. 


Fan 

1 


Fan 

2 


Hours 
Run. 


Gas 

Used. 

Pounds. 


Ice 

Made, 

Pounds. 


Engine 
Room. 


Boiler 
Room. 


Miscel- 
laneous. 



In the original copy all of these items are conveniently grouped on one large form ruled for 31-day entries 
with space at bottom for total quantities and costs. In the reproduction only the headings are included. 



FINANCE AND ECONOMICS — COST OF POWER 



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STEAM POWER PLANT ENGINEERING 









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FINANCE AND ECONOMICS — COST OF POWER 857 

TABLE 145. 

TYPICAL WEEKLY OPERATING CHART. 

The Edison Illuminating Company. 

Delray Power Houses, Detroit, Michigan. 

Date Dec. 5-12 inc., 1914. 

Pounds of coal per kilowatt-hour delivered 1.847 

B.t.u. per kilowatt-hour delivered 25,102 

Overall thermal efficiency, entire plant 13.59 

Output: 

Generated kw-hr. 6,860,900 

House Service, kw-hr. 170,500 

Delivered kw-hr. 6,753,400 

House Service to Total Generated Per Cent 1.574 

Average Output per Day kw-hr. 954,586 

Coal: 

Coal in Bunkers, Midnight, Dec. 5, lb. 15,748,400 

Coal to Bunkers, Dec. 5-12 Inclusive lb. 11,819,700 

Coal Chargeable, Dec. 5-12 Inclusive lb. 27,568,100 

Coal in Bunkers, Midnight, Dec. 12th lb. 15,094,000 

Coal Consumed, Dec. 5-12 Inclusive lb. 12,474,100 

Average Coal Consumed per Day, tons 891 

Coal Analysis: 

Total Moisture Per Cent 2.953 

Ash Per Cent 8.190 

Heating Value, As Fired (14004 dry) B.t.u. 13,592 

Ash Analysis: 

Carbon in Ash P. H. No. 1, Stokers with periodic 

dumping. . .Per Cent 20.165 

Carbon in Ash P. H. No. 1, Stokers with cinder 

grinders Per Cent 10.770 

Carbon in Ash P. H. No. 2, Stokers with periodic 

dumping ... Per Cent 21.775 

Carbon in Ash P. H. No. 2, Stokers with cinder 

grinders. . . . Per Cent 9.775 

423. Cost of Power. General. — The actual cost of producing power 
depends upon the geographical location of the plant, cost of fuel and 
labor, the size of apparatus, the design, conditions of loading system, 
of distribution and the method of accounting. Comparisons based on 
the cost per hp-hr. or per hp-yr., or the equivalent are without pur- 
pose because of the many variables entering into the problem. It is 
impossible to intelligently compare costs or to obtain a true under- 
standing of what costs for power really mean without a thorough 
knowledge of the various items entering into the unit cost such as 
costs of fuel, oil, waste, repairs, labor, insurance, taxes, management, 
distribution, maintenance and allowance for depreciation. In addition 
to these an understanding must be had of the operating conditions, 
such as size of plant, load factor, variation in load, ratio of the maxi- 
mum load to the economic full load, number of hours a day the plant 



858 STEAM POWER PLANT ENGINEERING 

is operated and the like. With each plant having an individuality 
distinctly its own, in so far as the charges which go to make up the 
ultimate cost is concerned, it is practically impossible to arrive at any 
definite conclusion as to the manner in which the real cost of power 
may be correctly determined for purposes of comparison. Perhaps 
the best method of stating station economy is to give the average 
yearly heat units supplied by the^ fuel per kw-hr. delivered to the 
switchboard, and the load factor. This eliminates price and quality 
of fuel and offers a fairly satisfactory criterion of the efficiency of 
operation. 

In any case the cost of power is based upon the expense which is 
independent of the output of fixed charges and that which is a function 
of the output or operating costs. In the small plant the items included 
in the fixed and operating costs are comparatively few in number and 
require but an elementary knowledge of bookkeeping, but in large 
industrial organizations or central stations the number of separate 
items to be considered may run' into the hundreds and necessitate a 
complex system of accounting. Some idea of the different systems 
employed with examples of cost of power in specific cases may be gained 
from an inspection of Tables 151 to 160. 

433. Fixed Charges. — These cover all expenses which do not expand 
and contract with the output. In the privately owned plant the fixed 
charges are usually limited to interest on the investment, rental, de- 
preciation, taxes, insurance and sometimes maintenance, though the 
latter is ordinarily included in the operating costs. The accounting 
systems for pubhc electric light and power companies are usually 
prescribed by the Public Utility Commission of the state in which 
the plant is located and the various charges must necessarily conform 
with the rules formulated by this Commission. 

In any system the total fixed charges per year are constant irre- 
spective of the load factor, since interest, taxes, depreciation, insurance, 
and maintenance go on whether the plant is in operation or not. The 
total fixed charges for a specific case are illustrated in Fig. 596 by a 
straight line. The cost per kilowatt-hour, however, decreases as the 
load factor increases. For example, with the plant operating con- 
tinuously at rated load (100 per cent load factor) the fixed charges 
per kilowatt-hour are 

With 30 per cent load factor these charges are 
65,000 



0.3 (5000 X 8760) 



00445 kilowatt-hour. 



FINANCE AND ECONOMICS — COST OF POWER 



859 



The higher the load factor the greater is the amount of power pro- 
duced and the longer does the apparatus work at best efficienc3^ But 
the greater the power produced the larger will be the fuel consumption 
and the oil and supply requirements. The labor charges will be prac- 
tically constant. The total operating cost per year increases as the 
load factor increases, but not directly. (See Fig. 596.) The cost per 



1.6 


























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i 


^ie.. 


^ 


a. 


^ 






40000 
































7 Cents 










00 


























1.,. 








1 








10 20 30 40 50 60 70 80 90 100 

Yearly Load Factor—Per Cent 

Fig. 596. Influence of Load Factor on the Cost of Power at the Switchboard. 
(5000-kilowatt Electric Light and Power Station.) 



kilowatt-hour, however, decreases as the load factor increases. For 
example, the operating costs per year with plant operating contin- 
uously at full load are $230,200. This gives 

230 200 
rr>^/^ w cnr. = $0.00525 pcr Idlowatt-hour. 
5000 X 8/60 



With 30 per cent load factor the yearly operating charges are $87,890, 
which gives 

87 980 
0.3 (5000 X 8760) = ^'^'^ P*"" kilowatt-hour. 

In general, the higher the load factor the greater becomes the ratio 
of the operating to the fixed charges, and extra investment may become 
advisable to secure the greatest economy possible. 



860 



STEAM POWER PLANT ENGINEERING 



On the other hand, when the load factor is low the fixed charges are 
the governing factor in the cost of power, and extra expenditures must 
be carefully considered, particularly if fuel is cheap. 

Fixed Costs in Industrial Power Plants : Engineering Digest, Apr., 1911, p. 293. 



TABLE 146. 

AVERAGE INITIAL COST.* 
Steam Engine Power Plants. 

Simple Non-Condensing. 



Horse Power. 


Dollars per 
Horse Power. 


Horse Power. 


Dollars per 
Horse Power. 


10 


225.00 


60 


180.00 


20 


200.00 


70 


177.00 


30 


195.00 


80 


175.00 


40 


190.00 


90 


170.00 


50 


185.00 


100 


.165.00 


Compound Condensing. 


100 


170.00 


700 


76.00 


200 


146.00 


800 


69.00 


300 


126.00 


900 


64.00 


400 


110.00 


1000 


60.00 


500 


96.00 


1500 


58.00 


600 


85.00 


2000 


55.00 


Triple Condensing. 


1000 


62.00 


4000 


52.00 


2000 


58.00 


5000 


50.00 


3000 


54.00 


6000 


48.00 - 



Includes cost of buildings and entire equipment erected. 



434. Interest. — The rates of interest on borrowed money vary with 
the nature of the security. In the case of power plants the form of 
security is usually a mortgage on the plant and equipment. If a 
builder has sufficient funds to construct the plant without borrowing, 
he should charge against the item ''interest" the income which the sum 
involved would bring if placed out at interest or if invested in his own 
business. In estimating the interest charges 6 per cent of the capital 
invested is ordinarily assumed unless specific figures are available. 
Initial costs for various types of plants are to be found in the accom- 
panying tables. 



FINANCE AND ECONOMICS — COST OF POWER 



861 



TABLE 147. 

COST OF MECHANICAL EQUIPMENT — ISOLATED STATIONS.* 



Boilers (erected and set in masonry): 

Horizontal-tubular 

Water-tube 

Steam engines: 

High-speed, simple direct-connected 

Medium-speed, compound non-condensing direct-connected. . . . 

Low-speed, compound condensing, belted 

Low-speed, simple, belted 

Gas engines 

Oil engines 

Gas producers 

Dynamos: 

Direct-connected to high-speed engine 

Belt-connected to engine 

Direct-connected to Corliss engine 

Switchboard 

Foundations 

Steamfitting — including auxiliary apparatus — such as feed heater. 

grease separator, exhaust head, tanks, covering, etc 



Per Kilowatt of 


Plant Capacity. 


$14-$18 


16- 20 


20- 


25 


28- 


35 


20- 


25 


25- 


30 


50- 


60 


75- 


85 


15- 


20 


13- 


16 


12- 


15 


16- 20 


5- 


10 


5- 


10 



20- 30 



* By P. R. Moses before the A.I.E.E., Jan. 12, 1912. 

435. Depreciation. — Depreciation may be defined as a decrease in 
value occasioned by wear or age, change of conditions rendering the 
plant inadequate for its particular functions, or changes in the art 
which renders it obsolete as compared with recent installations. De- 
preciation may be conveniently classified as: 

Complete depreciation, or the gradual decrease in value occasioned by 
wear and age. This. may be largely offset by maintenance. 

Obsolescence, inadequacy or destruction by any cause. A thing is 
obsolete when it has been rendered valueless as the result of change 
in the art and this may occur where no physical deterioration has 
taken place. Inadequacy indicates that a thing is incapable of fully 
performing the function for which it is intended. It indicates neither 
physical depreciation nor obsolescence. Inadequacy may result from 
expansion of markets, community growth and the like. Obsolescence, 
inadequacy, and destruction cannot be predicted and charges against 
this class of depreciation are naturally conjectural. 

Incomplete depreciation due to wear and tear likely to fall in large 
amounts and at irregular intervals. 

There are several methods of dealing with depreciation; among the 
more common may be mentioned : * 

* Report of the Committee on Gas, Oil, and Electric Light, City of Chicago, May, 
1913. 



862 



STEAM POWER PLANT ENGINEERING 



(1) To charge to earnings in good years and credit to depreciation 
reserve such amounts as the profits from operation permit. 

(2) To charge to earnings the depreciation as it matures and neces- 
sitates renewals. 

(3) To charge to earnings and credit to depreciation reserve an- 
nually a certain percentage of the cost determined by the average 
weighted life of the property. 

In general power plant practice it is customary to make an average 
annual depreciation allowance, based on the original cost of the property 
less salvage or junk value, spread over a period of years approximating 
the weighted life of the plant. If depreciation is considered to include 



TABLE 148. 

COST OF MECHANICAL EQUIPMENT — STEAM TURBO-ELECTRIC GENERATING 

STATIONS.' 
2,000 to 20,000-kilowatt Capacity, Based on Maximum Continuous Capacity of Generators at 50° Rise. 



Preparing site — Dismantling and removing structures from 
site, making construction roads, tracks, etc 

Yard Work — Intake and discharge flumes for condensing 
water, railway siding, grading, fencing sidewalks 

Foundations — Including foundations for building, stacks, and 
machinery, together with excavation, piling, waterproofing, 
etc 

Building — Including frame, walls, floors, roofs, windows and 
doors, coal bunker, etc., but exclusive of foundations, heat- 
ing, plumbing, and lighting 

Boiler-room Equipment — Including boilers, stokers, flues, 
stacks, feed pumps, feed-water heater, economizers, me- 
chanical draft, and all piping and pipe covering for entire 
station except condenser water piping 

Turbine-room Equipment — Including steam turbines and 
generators, condensers with condenser auxiliaries and water 
piping, oiling system, etc 

Electrical Switching Equipment — Including exciters of all 
kinds, masonry switch structure with all switchboards, 
switches, instruments, etc., and all wiring except for build- 
ing lighting 

Service Equipment — Such as cranes, lighting, heating, 
plumbing, fire protection, compressed air, furniture, per- 
manent tools, coal- and ash-handling machinery, etc 

Starting Up — Labor, fuel, and supplies for getting plant ready 
to carry useful load 

General Charges — Such as engineering, purchasing, super- 
vision, clerical work, construction, plant and supplies, 
watchmen, cleaning up 

Total cost of plant to owner, except land and interest during 
construction 



Dollars per Kilowatt. 



High. 



$0.25 
2.50 

6.00 

12.00 

24.00 
22.00 

5.00 

5.00 
1.00 

6.00 



$83.75 



Low. 



$.... 

1.00 
1.00 
4.00 

12.00 
12.00 

2.00 

2.50 
.50 

3.00 



$38.00 



• By O. S. Lyford, Jr., and R. W. Stoval, of Westing house, Church, Kerr & Company, before the 
Engineer's Society of Western Pennsylvania. 



FINANCE AND ECONOMICS — COST OF POWER 863 

the maintenance which is charged to expense directly, it would be proper 
to set aside as a reserve a fixed percentage of the decreasing value of 
the plant to represent the unmatured decadence. This ideal situation 
would equalize the total burden over the life by making the depreciation 
allowance largest when the repairs are smallest, and conversely the 
depreciation allowance smallest when the repairs are largest at the end 
of the useful life of the plant. If the system were composed of many 
small units not requiring renewal at or near the same time no special 
reserve would be necessary, as all replacements could be charged di- 
rectly to operating expenses because of these inconsiderable amounts 
in any one year. In the large central station, however, a considerable 
portion of the plant is composed of large units which the rapid develop- 
ment of the art and growth of business may render inadequate long 
before their natural life has expired. As a result of and to provide 
for this condition, depreciation reserves are accumulated either on the 
''straight line" or "sinking fund" method. 

Straight-line Method. — This method is based on the assumption 
that if the total investment, less salvage, is divided by the weighted 
life of the plant the resulting quotient expresses the amount which should 
be allowed each year to cover the accrued depreciation. This is the 
simplest of the several methods that have been suggested to determine 
the probable depreciation and make proper allowance for it in the rec- 
ords. No interest computations of any kind are involved, thus: 

r — SI 

D = - -, (304) 

n 

V = {C - S) (l - j\ (304a) 

(304b) 



(304c) 
(304d) 
(304e) 







d = 


100 


D 






V = 


100 


V 






A = 


c- 


-V, 






a = 


100 


A 


in which 










D = 


accrued depreciation. 




C = 


original cost, 








S = 


scrap value of salvage, 




n = 


assumed life, 


years, 




V = 


present value 


'; 







864 



STEAM POWER PLANT ENGINEERING 



m = age of the property, years, 

d = rate of depreciation, per cent of original cost, 

b = ratio of scrap value to original cost, 

V = present value, per cent of original cost, 
A = accrued depreciation, 

a = accrued depreciation, per cent of original value. 

The straight-line law is shown graphically in Fig. 597. The original 
cost is composed of the net cost (labor and material) plus the over- 
head (the extra charge intended to cover engineering and architects' 
fees; fire and liability insurance, and interest on the investment during 




Fig. 597. Straight-line Method of Depreciation. 

construction; contractors' profits on the portion of the work not done 
by the company itself; legal organization and incidental expense. The 
'^ probable life" of the plant is a purely theoretical quantity and is 
supposed to represent the weighted average period of usefulness of 
the various units composing the plant). It is determined by dividing 
the sum of the original costs of the various units composing the plant, 
less salvage, by the aggregate annual depreciation charge of these items. 
The actual life of the various units composing the plant can only be 
approximated since everything depends on the grade of material, work- 
manship and upkeep. Table 149 gives the average useful life of various 
portions of a steam power plant equipment, but so much depends upon 
the design and the conditions of operation that no fixed values can be 
definitely assigned and the values given should be used with caution. 
Most power plant appliances become obsolete long before the hmit of 
their useful life is reached. 

In the Report (May, 1913) to the Committee on Gas, Oil, and Electric 
Light on the investigation of the Commonwealth Edison Company, 
Chicago, depreciation is calculated on a 3 per cent sinking fund basis, 
giving an average weighted life of 18.5 years to the company's depre- 
ciable property. 



I 



FINANCE AND ECONOMICS — COST OF POWER 865 

TABLE 149. 

APPROXIMATE USEFUL LIFE OF VARIOUS PORTIONS OF STEAM POWER PLANT 

EQUIPMENTS. 

Years. 

Buildings, brick or concrete 40 

Buildings, wooden or sheet iron 15 

Chimneys, brick 40 

Chimneys, self-sustaining steel 30 

Chimneys, guyed sheet-iron 10 

Boilers, water-tube 25 

Boilers, fire-tube 20 

Engines, slow-speed 25 

Engines, high-speed 20 

Turbines 25 

Generators, direct-current 20 

Generators, alternating- current 20 

Motors 20 

Pumps 20 

Condensers, jet 25 

Condensers, surface 20 

Heaters, open 25 

Heaters, closed 20 

Economizers 20 

Wiring 15 

Belts 8 

Coal convf^yor, bucket 15 

Coal conveyor, belt 10 

Transformers 20 

Rotary converters 20 

Storage batteries 15 

Piping, ordinary 12 

Piping, first class 20 

Note. — So much depends upon the design and the conditions of operation that 
no fixed values can be definitely assigned and the above figures should be used with 
caution. Practice shows that most power-plant appliances become obsolete long 
before the limit of their useful life is reached. 

Example 75. A condenser equipment is 10 years old and cost origi- 
nally $3500.00. Assuming that its useful life is 25 years and that its 
junk value is $350.00, determine the annual depreciation, the present 
value and the accrued depreciation on the straight line basis. 

^ C - S 3500 - 350 ^^_ ,, . ^. , 

D = = ^^ = $126, annual depreciation charge. 

d = 100^ = 100 ^ = 3.5 per cent. 



V = {C - S) 



^ - ^) = (3500 - 350) (l- ^ = $1890, present 



value. 

V 1890 

V = 100 ^ = 100 ^^TTT^ = 54 per cent. 

A = C - V = $3500 - $1890 = $1610, accrued depreciation. 



866 STEAM POWER PLANT ENGINEERING 

Sinking Fund Method. — By the sinking fund method a fixed sum is 
placed aside each period and allowed to accumulate at compound in- 
terest. The amounts thus set aside plus the interest accumulations must 
be equal to the original cost less salvage at the end of the assumed period. 
The rate of depreciation in terms of interest and useful life is a simple 
problem in compound interest and may be expressed 

d = 100 y j^^ 7n^\ ' (305) 



(1 -\-ry - 1 
100' 

(1+r)- 



D = ^,, (30oa) 



a = 100 )\,\^ ^, (305b) 



A = ^{C-S), (305c) 

V = C -A, (305d) 

i;=100^, (305e) 

in which 

r = rate of interest, 

a = accrued depreciation, per cent of original cost less salvage. 

Other notations as previously designated. 

Example 76. Taking the data in Example 75 determine the annual 
depreciation charge, accrued depreciation, and present value on the 
sinking fund basis, assuming an annual interest rate of 5 per cent. 

J .r.r. r{l-h) ,^^ 0.05(1-0.1) 

d = 100 (i|,)._\ = 100 (1 ^ (;.Q5). _ 1 = 1-97 per cent. 

D = 0.0197 X $3500 = $68.95, annual payment to the sinking 
fund which at the end of 25 years will equal $3500 — $350 
= $3150. 

« = 100 [i ;,;._! - 100 I I ,,,;.. _ , = 26.35 per cent. 

A = $3150 X 0.2635 = 830.15 

V = $3500.00 - $830.15 = $2669.15, present value. 

z; = 100 ' ' =76.4 present value, per cent of original cost. 
ooUu 

Table 150 may be conveniently used in this connection. At the in- 
tersection of vertical column 5 and horizontal columns 10 and 25 we 
find 7.95 and 2.09 respectively. Dividing 2.09 by 7.95 gives 0.2635 
or 26.35 per cent, the accrued depreciation. 



FINANCE AND ECONOMICS — COST OF POWER 867 



TABLE 150. 



RATE OF DEPRECIATION. 



(Per Cent of First Cost.) 







Rate of Interest, per Cent. 




2 


2.5 


3 


8.5 


4 


4.5 


5 


5.5 


6 


7 


8 


9 


10 




2 


49.50 


49.37 


49.27 


49.14 


49.02 


48.90 


48.78 


48.66 


48.54 


48.31 


48.07 


47.84 


47.62 




3 


32.67 


32.51 


32.35 


32.19 


32.03 


31.87 


31.72 


31.56 


31.41 


31.10 


30.80 


30.51 


30.21 




4 


24.26 


24.08 


23.90 


23.72 


23.55 


23.39 


23.20 


23.03 


22.86 


22.52 


22.19 


21.84 


21.55 




5 


19.21 


19.02 


18.83 


18.65 


18.46 


18.28 


18.10 


17.91 


17.73 


17.40 


17.04 


16.73 


16.37 




6 


15.85 


15.65 


15.46 


15.26 


15.08 


14.89 


14.70 


14.52 


14.33 


13.97 


13.63 


13.29 


12.96 


5^ 


7 


13.45 


13.25 


13.05 


12.85 


12.66 


12.46 


12.28 


12.09 


11.91 


11.15 


11.20 


10.87 


10.55 


1 


8 


11.65 


11.44 


11.24 


11.05 


10.85 


10.66 


10.47 


10.28 


10.10 


9.74 


9.40 


9.06 


8.74 


ei 


9 


10.25 


10.04 


9.84 


9.64 


9.45 


9.26 


9.07 


8.88 


8.70 


8.34 


8.00 


7.68 


7.36 


§ 


10 


9.13 


8.92 


8.72 


8.52 


8.33 


8.14 


7.95 


7.76 


7.58 


7.23 


6.90 


6.58 


6.27 


< 

•4-1 


11 


8.21 


8.01 


7.80 


7.61 


7.41 


7.22 


7.04 


6.85 


6.68 


6.33 


6.00 


5.69 


5.40 


o 


12 


7.45 


7.25 


7.04 


6.85 


6.65 


6.46 


6.28 


6.10 


5.92 


5.60 


5.27 


4.97 


4.69 




13 


6.81 


6.60 


6.40 


6.20 


6.01 


5.83 


5.64 


5.47 


5.29 


4.96 


4.65 


4.36 


4.08 


3 


14 


6.26 


6.05 


5.85 


5.65 


5.46 


5.28 


5.10 


4.93 


4.75 


4.49 


4.13 


3.84 


3.58 


1 


15 


5.78 


5.57 


5.37 


5.18 


4.99 


4.81 


4.63 


4.46 


4.29 


3.97 


3.68 


3.40 


3.15 


P 


16 


5.36 


5.16 


4.96 


4.77 


4.58 


4.40 


4.22 


4.06 


3.89 


3.58 


3.30 


3.03 


2.78 


'v 


17 


4.99 


4.79 


4.59 


4.40 


4.22 


4.04 


3.87 


3.70 


3.54 


3.24 


2.96 


2.71 


2.47 


1 


18 


4.67 


4.46 


4.27 


4.08 


3.90 


3.72 


3.55 


3.39 


3.23 


2.94 


2.66 


2.42 


2.19 


1 


19 


4.37 


4.17 


3.98 


3.79 


3.61 


3.44 


3.27 


3.11 


2.96 


2.67 


2.47 


2.17 


1.95 


20 


4.11 


3.91 


3.72 


3.53 


3.36 


3.19 


3.02 


2.87 


2.71 


2.44 


2.18 


1.95 


1.95 




25 


3.12 


2.92 


2.74 


2.56 


2.40 


2.24 


2.09 


1.95 


1.82 


1.58 


1.36 


1.18 


1.75 




30 


2.46 


2.27 


2.10 


1.93 


1.78 


1.64 


1.50 


1.38 


1.26 


1.06 


0.88 


0.73 


0.61 




35 


2.00 


1.82 


1.65 


1.50 


1.36 


1.23 


1.10 


0.99 


0.89 


0.72 


0.58 


0.46 


0.37 




40 


1.65 


1.48 


1.32 


1.18 


1.05 


0.93 


0.83 


0.73 


0.64 


0.50 


0.38 


0.29 


0.22 




45 


1.39 


1.22 


1.07 


0.94 


0.82 


0.72 


0.62 


0.54 


0.47 


0.35 


0.26 


0.19 


0.14 




50 


1.18 


1.02 


0.88 


0.76 


0.65 


0.56 


0.42 


0.40 


0.34 


0.25 


0.17 


0.12 


0.09 



It is not supposed that an owner will regularly lay aside an annual 
amount, or take the trouble to arrange for its investment at current 
rates in the market or savings bank, since the money is probably worth 
more to him in his business. In practice it is retained in his business or 
investments and is earning the rate of interest obtainable therein, but in 
determining the net profit or loss this depreciation item is nevertheless 
accounted for just as if it were actually placed in outside investments. 

The expectancy or remaining life of any article is the probable time 
during which it may reasonably be expected to render efficient service. 
It is determined from the actual condition of the article and all local 
circumstances which may affect its continued use and not by subtracting 
age from probable life. Thus an article may have a probable life of 
25 years and yet be in first-class condition and as good as new when it 
reaches the end of this term. The value of this article is not written 
off the books nor should it be regarded as good as new. Its value is 
ascertained by determining its probable additional years of usefulness 
and the probable cost of replacing it at the end of this term. 



868 



STEAM POWER PLANT ENGINEERING 



The term ''depreciation" is frequently used when the term ''amortiza- 
tion" would be more appropriate. Amortization deals with the retire- 
ment of the invested capital. This may be in instalments in uniform 
or in unequal annual amounts, or in a lump sum at the end of useful 
life. The replacement may mean the substitution of a new identical 
plant, but at a cost dependent on new conditions, new prices of labor 




Fig. 598. Sinking Fund Method of Depreciation. 

and material, or it may mean the substitution of new devices rendering 
equivalent service. In either event the replacement may be at a greater 
or less cost than the original cost, with, therefore, a corresponding in- 
crease or decrease of capital invested. Expenditures for new parts of 
a plant, which take the place of old parts which are retired for any cause, 
should be charged to replacement only to the extent of capital repre- 
sented by the part of the plant thus retired. Any excess of the ex- 
penditure for replacement over the cost of the discarded part of a plant 
should be treated as an addition to, and any less cost as a deduction 
from, the invested capital. The term "replacement" should not be 
used in the sense of retirement of invested capital, which deals with the 
cost of the replaced part and not with the cost of the new equivalent 
installation. (Valuation Depreciation and the Rate-Base, Grunsky, 
1917.) 

The term going value may be properly taken to mean a value at- 
taching to a public utility property as the result of its having an estab- 
lished revenue-producing business. Going value may be determined 
from a consideration of the amounts of money actuall}' expended in 
the cost of producing the business or it may be determined from con- 
sideration of the present cost of reproducing the present revenue. 
(Value for Rate-Making, Floy, 1917.) 

For purpose of design and comparison it is customary to assume a 
single fixed percentage for depreciation, obsolescence, inadequacy, etc. 
An average figure is 5 per cent. 



FINANCE AND ECONOMICS — COST OF POWER 869 

426. Maintenance. — Maintenance usually refers to the expense of 
keeping the plant in running order over and above the cost of attend- 
ance, although the term is frequently used in place of ''repairs." It 
includes cost of upkeep, replacement, and precautionary measures. This 
latter item includes the renewal of working parts, painting of perishable 
or exposed material, and replacing worn-out and defective material. 
Many engineers make no allowance for maintenance in the fixed charges 
and include these costs under supplies, attendance, or repairs. In a 
general way, when maintenance is included under the fixed charges, an 
annual charge of 2 per cent is considered a liberal allowance, since most 
of the repair work comes under attendance. In street-railway practice 
maintenance is divided among the several parts of the system as follows: 
Buildings, steam appHances, electrical equipment, and miscellaneous. 
In this connection the maintenance becomes a part of the operating 
charges, since the various items vary widely from month to month. 

427. Taxes and Insurance. — Taxes vary from a fraction of one per 
cent to 2 per cent, depending upon the location of the plant. An aver- 
age figure is 1 J per cent of the actual value of the investment. Buildings 
and machinery are ordinarily insured against fire loss and boilers against 
accidental explosions, and accident poHcies are sometimes carried on all 
operating machinery. A fair charge for this item is one-half per cent. 

428. Operating Costs. — General Division. — The distribution of the 
operating costs depends largely upon the size and nature of the plant. 
In the small isolated station the term "operating costs" without qualifi- 
cation refers to the generating or station operating costs, exclusive of 
fixed charges. These costs are commonly divided as follows : 

1. Labor and attendance. 

2. Fuel and water. 

3. Oil, waste, and supplies. 

4. Repairs and maintenance. 

In some of the larger isolated stations a more extensive division is 
often made but there appears to be no accepted standard. 

In large central stations the operating costs are divided under the 
major headings of 

1. Production expenses. 

2. Transmission expenses. 

3. Electric storage expenses. 

4. Utilization expenses. 

5. Commercial expenses. 

6. New business expenses. 

7. General and miscellaneous expenses. 



870 



STEAM POWER PLANT ENGINEERING 



The extent of the subdivisions under each subheading depending 
upon the size and nature of the plant. See Table 151. 



TABLE 15L 

TOTAL EXPENSE (EXCLUSIVE OF DEPRECL\TION) FOR THE CALENDAR YEAR 1912. 
Commonwealth-Edison Co., Chicago. 





Total 
Expense. 


Per Cent 
of Total. 


798,677,000 
Kw-hr. Pur- 
chased and 
Generated. 
Cost in Cents 
Per Kw-hr. 


Production: 
Station wages 


$352,053 

2,137,076 

54,155 

37,887 

256,552 

359,311 


4.34 
26.38 
0.67 
0.47 
3.17 
4.44 


044080 


Fuel expense, including storage and shrinka"'e 


267577 


Station supplies and expense . 


0.006781 


Buildinf and property maintenance 


004744 




0.032122 


Purchased power 


0.044988 


Total production 

Transmission and distribution: 
Meter department expense . .... 


$3,197,034 

$237,090 

210,107 

110,088 

429,871 

60,503 


39.47 

2.92 
2.59 
1.36 
5.31 
75 


0.400291 
029685 




0.026307 


Storage batterv operating and repairs 


013784 


Maintenance of overhead and undero^round lines 


053823 


Install, remove, exchange meters 


0.007575 


Total transmission and distribution 


$1,047,659 

$224,538 

77,618 

345,465 

43.364 


12.93 

2.77 

0.96 
4.26 
0.54 


0.131174 


Utilization: 
Maintenance tungsten fixtures and posts ) 


0.028114 


Maintenance arc lights ) 

Repairs to customers' installations 


0.009718 




0.043255 


Inspection of customers' premises 


0.005429 


Total utilization 


$690,985 

$209,669 

211,635 

61,042 


8.53 

2.. 59 
2.61 
0.75 


086516 


New business: 
Contract department expense ... 


026252 


Advertising 


0.026498 


Wiring and appliances 


0.007643 


Total new business 


$482,346 

$152,007 
11,025 
9,072 
68,581 
22.206 


5.95 

1.88 
0.14 
0.11 
0.85 
0.27 


060393 


Commercial expense: 
Collecting and bookkeeping 


019032 


Claim department expense 


0.001380 




0.001136 


Billing department expense 

Customers' statistics 


0.008587 
0.002780 


Total commercial expense 


$262,891 

$483,717 
152,633 
51,069 
89,954 
179,618 
103,055 
69,850 


3.25 

5.97 
1.88 
0.63 
1.11 
2.22 
1.27 
86 


0.032916 


General expense: 

Executive and legal and loss and damage account 

Maintenance and rental of offices and miscellaneous buildings . . 
Telephone and telegraph and general office sundries 


0.060565 
0.019111 
0.006394 


Purchasing and stores department expense 


011263 


Engineering and operating supervision 


0.022489 


General office departments, accounting and statistics 


012903 


Net profit on mercantile sales 


0.008746 


Total general expense 


$990,196 

$22,627 

90,509 

5,436 

460,195 

714,000 

277,017 

117,225 

193,316 

72,000 

39,114 

70.077 


12.22 

0.28 
1.12 
0.07 
5.68 
8.81 
3.42 
1.45 
2.38 
0.89 
0.48 
0.87 


123980 


Miscellaneous: 

Transportation department undistributed and miscellaneous 

Miscellaneous operating — steam . . 


0.002833 
011332 


Conduit rental 


000681 


Municipal compensation 


057620 


Taxes 


0.089398 


Insurance 


034684 


Interest, discount, and exchange 


0.014677 


Profit on .stores 


024205 


Pension fund 


009015 


Discount on bonds 


0.004897 


Bad debts 


0.008774 


Total miscellaneous expense 


$1,429,562 
$8,100,673 


17.65 


178991 






Grand total 


100.00 


1.014261 







FINANCE AND ECONOMICS — COST OF POWER 871 

A number of large central stations limit the major headings to 

1. Generation. 

2. Administration. 

3. Distribution. 

Some companies include all or part of the fixed charges under the 
major heading, others limit the operating costs to expense which is 
dependent only on the output. Because of this diversity in book- 
keeping comparisons of the cost of power based on the annual report are 
without purpose. An excellent system is that prescribed by the State 
Board of Pubhc Utility Commissioners of New Jersey, a discussion of 
which is to be found in Power, Nov. 11, 1913, p. 697. A few annual 
reports illustrating the different systems of accounting are reproduced 
in the accompanying tables. 

439. Labor, Attendance, Wages. — The minimum number of men 
required to handle a given plant is approximately a fixed quantity and 
it is seldom possible to so arrange the work that any material reduction 
can be effected. Until very recently it has been the universal custom 
to pay wages on a ''flat rate" basis, that is, the attendant is given a 
fixed sum per day or month irrespective of the amount of work required 
or the economy of operation. In many cases, however, the bonus 
system has been successfully adopted. For example, in the boiler room 
the coal consumption is determined for a given period of time with 
ordinary careful firing, and the fireman is offered a reasonable per- 
centage on the saving of coal which he is able to effect over this record 
by special care and attention to the keeping of fires always in the best 
condition, avoiding the blowing off of steam, using as little coal as 
needed for banking fires, and in other ways. Where careful records are 
kept of supplies, repairs, and renewals, the bonus is also apphcable to 
electricians, oilers, and other employees. 

Labor should always be estimated or recorded as so many dollars 
per month or per year and not merely in terms of the output unless 
the load factor is definitely known, otherwise comparisons are mis- 
leading. For example, consider two plants of 500 kilowatts capacity, 
each with labor charges, say, of $400 per month. Suppose the output 
of one is 100,000 kilowatt-hours per month and that of the other 40,000 
kilowatt-hours per month. The monthly charges are evidently the 
same, viz., $400, but the cost per kilowatt-hour differs widely, being 
0.4 cent in the first case and 1 cent in the latter. 

The cost of labor varies so much with the location of the plant and the 
conditions of operation that general figures are of little value except as a 
rough guide. Specific figures will be found in the accompanying tables. 

For a summary of labor costs in largo central stations see " Central-Station Labor 
Costs," Electrical World, Nov. 16, 1912, p. 1031. 



872 



STEAM POWER PLANT ENGINEERING 



430. Cost of Fuel. Tables 151 to 160 give specific examples of 
the cost of fuel in different sizes and types of steam power plants. 
It will be noted that this item varies considerably even with plants 
of the same general class. So much depends upon the grade and market 
price of the fuel, type, and size of plant and conditions of operation that 
no single item can afford a means of comparing fuel costs in different 
plants. Such items as ''lb. coal per kw-hr., " " cost of fuel per kw-hr., " 
or the equivalent have their value in any accounting system, but fail 
utterly as a measure of the economy of operation unless accompanied 
by a statement of the qualifying conditions. For example, an ineffi- 
ciently operated plant using a high-grade fuel may show a lower fuel 
consumption, lb. per kw-hr., than an economical plant using a low- 
grade fuel, and an uneconomical plant using a very cheap fuel may show 
a lower ''cost of fuel per kw-hr." than an efficiently operated plant 
using costly fuel. Similarly, two plants of the same size and type, and 



TABLE 152. 

FUEL CONSUMPTION IN MASSACHUSETTS CENTRAL STATIONS. 

(Year ending 1915.) 



Company. 


Long Tons Used. 


Cost per 
Ton. 


Total Coal 
Cost. 


Cents 
Per 

Kw-hr. 

Gener- 
ated. 


Lb. 

Coal 

Per 

Kw-hr. 


Cambridge El. Lt. Co 

Easthampton Gas Co. 


15.251 
j 4,170 
) 2.35 coke 
182,679 
15,625 
14,871 
( 8,436 
] 4,328 coke 
( 348 gas coal 
4,494 
9,600 
( 4,574 
] 39 coke 
( 1,561 dust 
18,.584 
16,589 
( 14,436 
1 1,884 coke 

10,935 

{ 7,532 

1 96 gas coal 

8,973 

31,954 

7,300 

37,462 


$4,022 

4.351 ( 

3.50 j 

3.902 

4.778 

3.673 

4.167) 

4.48 > 

4.55 ) 

4.623 

4.664 

4.667) 

4.000 > 

2.000) 

4.698 

3.545 

4.6 I 

4.0 f 

3.608 

4.115) 

5.6 ( 

4.074 

4.277 

4.534 

4.142 


$61,339 

18,150 

712,734 
74,660 
54,621 

56,127 

20,778 

44,777 

24,627 

87,316 
58,799 

73,951 

39,457 

31,531 

36,557 
136,686 

33,096 
155,180 


403 

0,641 

0.359 
0.458 
0.390 

0.723 

0.693 
0.64 

0.995 

0.667 
0.461 

0.693 

0.472 

0.618 

0.55 
0.567 
0.579 
0.48 


2.246 
3 301 


Edison Elec, 111., Boston 

Edison Co., Brockton 


2.063 
2 149 


Fall River El. Lt. Co 


2 378 


Fitchburg Gas & El. Co 

Greenfield El. Lt. Co. 


3.785 
3 359 


Haverhill Electric Co. 


3 075 


Lawrence Gas Co 


5 588 


Lowell El. Lt. Co 


3 178 


Lynn Gas & Electric Co 

Maiden Electric Co 


2.913 
3 424 


New Bedford Gas & El. Lt 

No. Adams Gas Lt. Co 

Salem El. Lt. Co 


2.931 
3.351 
3 023 


Springfield United El. Lt 

Webster & S. Gas & El. Lt 

Worcester El. Lt. Co 


2.971 
2.86 
2 593 







The Cambridge, Boston, Fall River, Lynn, New Bedford, and Salem companies are located on tide- 
water and enjoy the advantage of cheaper fuel transportation than those located inland. 



FINANCE AND ECONOMICS — COST OF POWER 



873 



TABLE 153. 

POWER COSTS IN CENTRAL STATIONS. 

Station A. 10-500 hp. boilers; 5000 hp. piston engines; 111. screenings; no 
coal-handling ai)paratus; hand-fired furnaces. 

Station B. Modern steam turbine plant; stoker equipment; coal- and ash- 
handling system; economizers; superheaters; 111. screenings. 

Station C. 5400 hp. boilers; 14,000-kw. turbines and engines; coal- and ash- 
handling system; stoker equipment; 111. screenings. 

June, 1913. 



Kw-hr. generated 

Tons of coal 

Tons of ash 

Lb. water evaporated 

Lb. water evaporated per lb. coal . 

Lb. coal per kw-hr 

Lb. water per kw-hr 

Gal. engine oil per 1000 kw-hr 

Gal. cylinder oil per 10,000 kw-hr. 



1,061,000 

2,775 

555 

40,600,000 

7.32 
5.23 



3.62 
1.74 



B. 



1,210,750 
2,437.37 
322.10 
35,359,500 

7.25 
4.03 
29.20 
0.59 
0.39 



1,404,605 
3,981.40 
603 
58,100,000 
7.5 
5.55 
41.6 
1.94 
1 22 



Total Cost, and Cost Per Kw-hr. in Cents. 



Kw-hr, 



Kw-hr, 



Kwhr. 



Superintendence 

Repairs: 

Dynamos and appliances 

Engines 

Boilers 

Pumps, pipes, fittings and miscellaneous. 

Operating boilers 

Operating engines and dynamos 

Supplies 

Water 

Lubricants and waste 

Miscellaneous expense 

Total, except fuel 

Coal 



122.42 
171.33 



Coal, labor, car to boiler room 

Total cost 

Average cost of coal per ton on floor of boiler room 



1017.48 

8.80 

880.92 

693.66 

5.47 

482.21 

220.12 

291.24 

3893.65 

2635.75 

198.62 



0.014 
0.019 

o.m 

0.001 
0.100 
0.079 



0.055 
0.025 
0.033 
0.441 
0.298 
0.022 



250.10 

10.84 

'299!8i 

22,15 

392.13 

390.00 

44.80 

99.75 

42.50 

60.08 

1612.16 

2177.44 

114.62 



0.020 

0.001 

'6;024 
0.002 
0.033 
0,032 
0.004 
0.008 
0.004 
0.005 
0.133 
0.180 
0.009 



246.47 
12.94 



1332.34 
794.96 
689.79 
101.10 

■ "95;74 

177.68 

3451.02 

4906.44 

105.48 



0.018 
0.001 



0.094 
0.057 
0.049 
0.007 

'6!667 

0.013 
0.246 
0.349 
0.008 



6728.02 
1.0214 



0.761 



3904.22 
0.94 



0.322 



8462.94 
1.344 



0. 



October, 1913. 



B. 



C. 



Kw-hr. generated 

Tons of coal 

Tons of ash 

Lb. water evaporated 

Lb. water evaporated per lb. coal . 

Lb. coal per kw-hr 

Lb. water per kw-hr 

Gal. engine oil per 10,000 kw-hr. . . 
Gal. cylinder oil per 10,000 kw-hr. 



1,356,610 

3,052.4 

610.5 

30,681,000 

6.5 

4.5 



1.24 
7.07 



1,215,360 
2,838.5 
456.53 
35,625,500 

6.28 
4,67 
29,31 
0,41 
0.41 



1,704,596 
4,900,72 
1,080 
64,484,866 

6,58 
5,75 
37.83 
0,95 
0.40 



Total Cost, and Cost per Kw-hr. in Cents. 







Kw-hr. 




Kw-hr. 




Kw-hr. 


Superintendence 

Repairs: 
Dynamos and appliances 


121.67 
245.18 


0.010 

0.020 

0:646 
0.001 
0.050 
0.059 
0.004 
0.029 
0.019 
0.007 


243.74 

21.93 

"484:5i 

9.00 

396.15 

390.00 

16.50 

98.16 

37.50 

41.89 


0.020 

0.002 

'6:040 
0,001 
0,033 
0,032 
0,001 
0.008 
0.003 
0.003 


201.14 

469.43 
66.78 
833.61 
595.97 
843.68 
673.29 
116.25 

"1.50:66 
246 18 


0.012 

0.028 
004 


Boilers . . . 


559.11 

16.32 

608.64 

718.88 

41.65 

354.97 

228.82 

78.39 


049 


Pumps, pipes, fittings and miscellaneous. .... . 

Operating boilers . . ... 


0.025 
049 




0.039 


Supplies 


007 


Water 




Lubricants and waste 

Miscellaneous expense 


0.009 
0.014 


Total, except fuel 


2973.63 

2899.91 

187.60 


0.245 
0.239 
0.015 


17.39.38 

2469.50 

135.60 


0.143 
203 
0.011 


4,196.99 

6,150 62 

183 20 


246 


Coal 

Coal labor, car to boiler room 


0.361 
0.011 


Total cost 


6061 14 
•SI. 01 13 


0.499 


4344,48 
$0.9178 


0.357 


10,530 81 
$1,255 


618 


Average cost of coal per ton on floor of boiler room 





874 STEAM POWER PLANT ENGINEERING 

using the same fuel may show considerable difference in both ''\h. of 
fuel per kw-hr. " and ^'cost of fuel per kw-hr. " because of difference 
in load factor even though both plants are efficiently operated for the 
given conditions. In a number of recent installations the station oper- 
ating records include the heat supplied by the fuel per kw-hr. generated 
C^B.t.u. per kw-hr.") and the cost of the fuel on a heat basis (cents per 
10,000 B.t.u.). These two items in connection with the load factor 
offer a satisfactory criterion of the fuel economy for plants of the same 
general design. Large central stations with individual units of 20,000 
to 35,000 kw. rated capacity and yearly load factor of 50 per cent or more, 
have been credited with a yearly performance of 20,000 B.t.u. per kw-hr. 
generated, corresponding to an overall thermal efficiency of 17 per cent. 
With Ilhnois screenings this is equivalent to approximately 2 lb. coal 
per kw-hr. and with the better grades of bituminous coal, about 1.5 lb. 
coal per kw-hr. Much better results than this have been obtained for 
brief periods of operation but when averaged over a considerable period 
of time the standby losses, such as coal burned in banking fires, heat 
lost in blowing down boilers, lower efficiency in operating at underloads 
and overloads and the hke, reduce the overall efficiency to substantially 
that given above. The coal consumption per kw-hr. for a number 
of medium size central stations in Massachusetts is given in Table 152. 
This table does not offer a fair basis of comparison since the calorific 
value of the fuel and the yearly load factor are not given. 

In estimating the cost of fuel for a proposed installation the logical 
procedure is as follows : 

1. Construct load curves for the probable power requirements. 

2. Calculate the total weight of steam suppUed from the load curve. 

3. Transfer the total steam requirements to the unit water rate basis. 

4. Reduce the average unit water rate to ''B.t.u. supplied by the 
steam per unit output." 

5. Divide the average B.t.u. suppfied by the steam per unit output 
by the estimated overall boiler efficiency, considering all standby loss. 
This gives the B.t.u. supplied by the fuel per unit output. 

6. Reduce the cost of fuel to ''cost per 10,000 B.t.u." 

7. Multiply item 5 by item 6 and divide by 10,000. This gives the 
average cost of fuel per unit output for the required period. 

The construction of the load curves is the most important item since 
the cost of the fuel per unit putput is primarily a function of the load 
factor. See paragraph 434. 

The total weight of steam is calculated from the load curve by con- 
sidering the unit water rate of the prime mover and steam-driven 
auxiliaries at the variable loads^ and the time element. 



FINANCE AND ECONOMICS — COST OF POWER 



875 



TABLE 154. 

DISTRIBUTION OF STATION OPERATING COSTS. 

Steam Turbine Plants. (Medium Size.) 

(Year Ending 1915.) 





Brockton. 


Fall River. 


New 
Bedford. 


I'nited 
Elec. Light. 


Rated boiler capacity, hp. . 


3300 


2800 


3416 


9300 


Rated turbine capacity. k\v. 


9000 


10,000 


9400 


13,600 


Output, k\y-hr. (million). . 


16.28 


14.00 


8.35 


24.09 


Load factor, per cent 




32.4 






Tons of coal (thousands) . . 


15.62 


14.87 


10.96 


32.0 


Coal per k\v-hr., lb 


2.12 


2.38 


2.93 


2.82 


Cost of coal per ton 


$4.78 


$3.67 


$3.60 


$4.27 


Men employed 


26 


20 













Operating Costs, Cents per 


Kw-hr. 












Actual. 


Per 
Cent 
Total. 


Actual. 


Per 

Cent 
Total. 


Actual. 


Per 
Cent 
Total. 


Actual. 


Per 
Cent 
Total. 


Fuel 


0.458 

0.005 
0.019 
0.179 

0.023 
0.079 
0.069 
0.020 


53.9 

0.6 

2.2 

21.0 

2.7 
9.2 
8.1 
2.3 


0.390 
0.005 
0.016 
0.124 

0.011 
0.011 
0.024 
0.015 


65.4 

0.8 

2.7 

20.8 

1.9 
1.9 
4.0 
2.5 


0.472 
0.003 
0.038 
0.309 

0.025 
0.017 
0.027 
0.005 


52.7 
0.3 
4.2 

34.5 

2.8 
1.9 
3.0 
0.6 


0.570 
0.005 
0.004 
0.160 

0.008 
0.050 
0.081 
0.033 


62.5 


Oil, waste, and packing . . . 
Water 


0.6 
0.5 


Wages 


17.6 


Station tools and appli- 
ances 


0.8 


Station structure repairs. . 

Steam plant repairs 

Electric plant repairs 


5.5 
8.9 
3.6 


Total 


0.852 


100.0 


0.596 


100.0 


0.896 


100.0 


0.911 


100.0 







TABLE 155. 

STATION OPERATING COSTS (1915). 
Massachusetts Steam Power Plants. 



Plant. 



Cambridge 

Easthampton 

Edison, Boston. . . 
Edison, Brockton. 

FallRiyer 

Haverhill 

Lowell 

Lynn 

Maiden 

New Bedford 

Salem 

Worcester 



Fuel. 


Oil, 
Waste 
and 
Pack- 
ing. 


Water. 


Wages. 


0.403 
0.641 
0.359 
0.458 
0.390 
0.640 
0.667 
0.461 
0.693 
0.472 
0.550 
0.480 


0.012 

0.004 
0.002 
0.005 
0.005 
0.016 
0.006 
0.015 
0.011 
0,003 
0.013 
0.004 


0.026 
0.003 

0.010 
0.019 
0.016 

0'007 
0.032 
0.057 
0.038 
0.020 
0.004 


0.274 
0.307 
0.161 
0.179 
0.124 
0.209 
0.193 
0.194 
0.177 
0.309 
0.233 
0.108 



Station 
Tools 
and 
Appli- 
ances. 



0.003 
0.005 

017 
. 023 
0.011 
0.011 
0.013 
0.003 
0.015 
0.025 
0.010 
0.003 



Station 
Struc- 
ture 
Re- 
pairs. 



0.017 
0.001 

0.009 
0.079 
0.011 
0.055 
0.032 
0.051 
0.002 
0.017 
0.002 
0.018 



Steam 
Plant 
Re- 
pairs. 



0.040 
0.049 
0.051 
0.069 
0.024 
0.071 
0.062 
0.152 
0.058 
0,027 
0.059 
0.046 



Elec- 
trical 
Station 
Re- 
pairs. 



0.046 

0,010 
0,060 
0,020 
0.015 
0,004 
0,005 
0.014 
0.006 
0,005 
0,006 
0,014 



Total. 



0,821 

1,020 
0,687 
0.852 
0.596 
1.006 
0.986 
0.922 
1.019 
0.896 
0.903 
0.680 



876 



STEAM POWER PLANT ENGINEERING 



The heat suppUed by the steam is measured above the temperature 
of the feed water. In plants where exhaust is used for heating or 
manufacturing purposes only the difference between the heat supphed 
to the prime movers and steam-driven auxiliaries and that of the 
exhaust utihzed for heating is charged to power. See paragraph 177. 

Current practice gives an average efficiency (based on yearly opera- 
tion) of boiler and furnace of 70 per cent for pumping stations running 
at practically full load, 68 per cent for large lighting and power stations 

50 




1890 1900 1910 

Fig. 599. Development of the Steam Power Plant. (Locomobile Type.) 

with yearly load factor of 0.45 or more, and 65 per cent for similar 
stations with load factor between 0.35 and 0.40. For very low load 
factors, 0.25 and under (as in connection with large manufacturing 
plants, tall office buildings, and other plants operating on a 12-hour 
basis), this efficiency seldom exceeds 60 per cent. With these figures 
as a guide the cost of fuel per unit output may be roughly approximated. 
In Europe the "locomobile" type of steam power plant has attained 
an extremely high degree of heat efficiency as will be seen from the 
curves in Fig. 599. The most economical result shown, namely 0,87 



FINANCE AND ECONOMICS — COST OF POWER 



877 



pound of coal per developed horsepower-hour, is equaled only by our 
best gas-producer plants. 

431. OU, Waste, and Supplies. — These items approximate from a 
fraction to 5 per cent of the total operating expenses. Tables 153 to 



TABLE 156. 

YEARLY COST OF OPERATION. 

Fort Wayne Municipal Plant. 

(1915-1916.) 



Equipment: 2-500, 1-1500, 1-300 = 5500 kw. turbo-generators. 
1-725, 2-500, 1-400, 3-300 = 3025 hp. boilers. 



Investment cost: 

Boiler-plant equipment 

Boiler-plant buildings, fixtures and grounds 

Steam power plant equipment 

Steam power plant building, fixtures and 
grounds 

Total power plant 

Distribution system and other expenses. . . . 

Grand total 

Total output 6,520,670 kw-hr. 

Total coal burned 18, 100 tons 

Yearly load factor 24.7 per cent 



Total. 



'$79,363.75 

26,302.34 

182,773.75 

39,469.54 



327,909.38 
407,138.19 



Unit. 



$26.00 per b 

8.70 " 

33.00 " kw. 

7.20 '* '' 

59.50 " '* 

74.00 " " 

134.00 " " 



hp. 



735,047.57 

Lb. coal per kw-hr. 



5.55 



Station operating costs: 

17 men, 3-8 hr. shifts, labor . 

Coal, $1 .80 per ton delivered 

Supplies and sundries 

Maintenance 

Total 

Total expense: 

Steam power generation 

Distribution 

Consumption 

Commercial 

General 

Depreciation 

Undistributed 

Contingencies 

Grand total 



Cost per Year. 



$17,296.11 

32,578.48 

514.71 

6,980.40 



$57,369.70 

$57,369.70 
10,269.01 
17,730.74 
11,472.04 
12,144.63 
29,658.99 
9,850.30 
4,391.98 



$152,887.39 



Per Cent 
of Total. 



30.2 

56.8 

0.8 

12.2 



100.00 

37.6 
6.7 

11.6 
7.5 
7.9 

19.5 
6.4 
2.8 



100.0 



Cents 
per Kw-hr. 



0.265 
0.499 
0.008 
0.107 



0.879 

0.879 
0.158 
0.272 
0.175 
0.185 
0.455 
0.149 
0.067 



2.340 



160 give some idea of current practice in different classes of power 
plants. 

433. Repairs and Maintenance. — This item ordinarily refers to the 
cost of keeping the plant in running order over and above the cost of 



878 



STEAM POWER PLANT ENGINEERING 



labor or attendance, and depends upon the age and condition of the 
plant and the efficiency of the employees. Tables 153 to 160 give the 
cost of repairs and maintenance for a wide range in power-plant 
practice. 

433. Cost of Power. — The actual cost of producing power depends 
upon the geographical location of the plant, the size of apparatus, the 
design, conditions of loading, system of distribution, and the method 
of accounting. Tables 151 to 160 compiled from various sources give 
the detailed costs of a large number of central and isolated stations. 



TABLE 157. 

COST OF GENERATING 1000 LB. STEAM. ■• 

N. Y. Buildings — Steam Heating Only. 

(1915.) 



No. of Building. 



Type of building 

No. of floors ■ 

Building vol. cu. ft. (million) 

Duration of test, days 

Steam generated, 1000 lb 

Tons of coal, gross 

Rate of evaporation 

Average outside temperature. , 

Boiler capacity, hp 

Maximum boiler, hp 

Average boiler, hp 



25 



15 

1611 

117 

6.15 

30.7 

384 

280 

100 



L 

12 

4 

4 

362 

27.6 

5.84 

34.2 

600 

300 



O 

25 

6.5 

5 

485 

30.3 

7.13 

39.6 

600 

330 

.150 



4 

124 

11.3 

4.94 

37.0 

800 

150 

50 



D 

'"l5 

46 

10,310 

783 

5.89 

34.8 

1200 

600 

235 



O 



151 

36,890 

2540 

6.31 

40.9 

900 

850 

350 



Cost per 1000 Lb. Steam. 



Coal 

Labor 

Ash removal 

Water (makeup) 

Electric current (forced draft) 

Electric current (boiler feed pump) 

Supplies 

Repairs and miscellaneous 

Total 

Fixed charge on investment 

Total cost per 1000 lb 



191 
049 
010 



0.014 



004 
004 



272 
029 
301 



10.201 
0.085 
0.011 

0'005 

0^011 
0.004 



$0,317 
0.051 



SO. 368 



$0,165 
0.079 
0.009 

0.007 
0.007 
0.006 
0.002 



$0,275 
0.054 



$0,329 



$0,238 
0.251 
0.021 

0.021 

0.002 
0.001 



$0,535 
0.084 



$0,619 



$0,203 
0.052 
0.008 
0.007 
0.008 

o'o64 

0.003 



$0,285 
0.044 



$0,329 



$0,187 
0.056 
0.007 
0.001 
0.006 

6;666 

J. 002 



$0,265 
0.033 



$0,298 



" O," Office building; " L," Loft building; " D," Department store. Coal, $2.50 per ton in all 
buildings. 

* From report of the Station Operating Committee, National District Heating Association, read 
June 3, 1915, at Chicago. 



FINANCE AND ECONOMICS — COST OF POWER 



879 



TABLE 158. 

SOME POWER COSTS FROM A MODERN APARTMENT HOUSE. 

(New York.) 

Original cost of plant on the foundation, 1909 $113,424 

Present value at 10 per cent charged off each year 60,279 

Average Cost per 24 hr. for 1916: 

Labor S39.59 

Coal 54.13 

Ashes 1 . 66 

Oil 1.27 

Supplies 7.01 

Repairs 4.61 

Improvements .65 

Depreciation (10 per cent on $60,279) 16.51 

Total cost per 24 hr $125 . 43 

Average cost per hr 5 . 22 + 

Quantities and Costs for Year Ended Dec. 31, 1916: 
Water consumed in boilers per 24 hr. (venturi-meter measured), lb. . . 376,911 

Coal consumed per 24 hr., lb 38,828 

Ashes put out per 24 hr., lb 6,538 

Average horsepower-hr. developed per 24 hr 10,925 . 04 

455.21 

9.271 

9.707 
3.55 
16.8 
12.48+ 
0.0114 



Average horsepower-hr. developed per hr 

Water evaporated per pound of coal (actual conditions), lb 
Water evaporated per pound of coal (from and at 212), lb. . 

Coal (No. 3 Buck.) consumed per hp.-hr., lb 

Ash, per cent 

Ash per analysis (commercial) 

Cost per hp-hr., dollars 



Kw-hr. dehvered to board for 1916 788,129 

2,159 

90 



Average kw-hr. per 24 hr 

Average kw-hr. per hr 

Electric load was 21 per cent of total load and 

Cost per 24 hr., dollars 

Cost per hr., dollars 

Cost per kilowatt-hours, dollars 

Income from store Ughting per 24 hr., dollars . 

Net operating cost per 24 hr., dollars 

Net operating cost per hr., dollars 

Net hp-hr. cost, dollars 

Net kw-hr. cost, dollars 



26.34 
1.10- 
0.012 + 
9.06 
116.37 
4.85 
0.0107 
0.0114 



Year. 


Average B.t.u. 


Average Ash, 
Per Cent. 


Average 
Moisture, 
Per Cent. 


Average Coal per 
Hp-hr.. Lb. 


Average Coal 
Cost per 
Hp-hr., 
Dollars. 


1912 

1913 

1914 

1915 

1916 


12,672.22 
12,538.46 
12,826.43 
12,825.01 
12,796.40 


14.70 

14.687 
14.425 
13.57 
12.48 


6.81 

7.05 
6.386 
6.75 
7.05 


3.891 No. 1 
4.427 Nos. 1, 2&3 
3.487 No. 3 
3.329 No. 3 
3.554 No. 3 


0.0059 

0.0060 
0.0051- 
0.0043 
0.0049+ 



880 



STEAM POWER PLANT ENGINEERING 



TABLE 159. 

COST OF ONE HORSE POWER PER YEAR, SIMPLE ENGINES, NON-CONDENSING. 

10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineering Magazine, July, 1908, p. 563.) 



Size of plant horse power 

Cost of plant per horse power 

Fixed charges at 14 per cent 

Coal per horse-power hour, in pounds. . . . 

Cost at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, and supplies 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



20 


40 


60 


$200.00 


$190.00 


$180.00 


28.00 


26.60 


25.20 


12.00 


10.00 


9.00 


66.00 


55.00 


49.50 


30.00 


20.00 


15.00 


6.00 


4.00 


3.00 


146.50 


119.35 


105.07 


138.25 


112.47 


98.80 


130.00 


105.60 


92.70 


121.75 


98.72 


86.51 


113.50 


91.85 


80.32 


105.25 


84.97 


74.13 


97.00 


78.10 


67.95 



80 
$175.00 
24.50 

8.00 
44.00 
13.00 

2.60 
95.10 
89.60 
84.10 
78.60 
73.10 
67.60 
62.10 



TABLE 160. 

COST OF ONE HORSE POWER PER YEAR, COMPOUND CONDENSING ENGINES, 

10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineering Magazine, July, 1908, p. 564.) 



Size of plant horse power 


100 


200 


300 


400 


500 


600 


Cost of plant per horse power. . . . 


S170.00 


$146.00 


$126.00 


$110.00 


$96.00 


$85.00 


Fixed c larges at 14 per cent 


23.80 


24.40 


17.65 


15.40 


13.45 


11.90 


Coal per horse-power hour, pounds 


7.0 


6.5 


6.0 


5.5 


5.0 


4.5 


Cost of fuel at S4.00 per ton 


38.50 


35.70 


33.00 


32.00 


27.50 


24.70 


Attendance, 10-hour basis 


12.00 


10.00 


8.60 


7.25 


6.20 


5.40 


Oil, waste, supplies 


2.40 


2.00 


1.72 


1.45 


1.24 


1.08 


Total 


76.70 


68.10 


60.97 


56.10 


48.39 


43.08 


With coal at $5.00 per ton 


86.40 


77.10 


69.22 


61.90 


55.29 


49.28 


With coal at $4.50 per ton 


81.50 


72.60 


65.07 


58.10 


51.79 


46.18 


With coal at $4.00 per ton 


76.70 


68.10 


60.97 


56.10 


48.39 


43.08 


With coal at $3.50 per ton 


71.90 


63.70 


56.82 


50.50 


45.04 


30.98 


With coal at $3.00 per ton 


67.00 


59.20 


51.67 


46.70 


41.49 


36.88 


With coal at $2.50 per ton 


62.30 


54.75 


48.59 


43.00 


38.83 


33.83 


With coal at $2.00 per ton 


57.45 


50.25 


44.47 


40.10 


34.64 


30. 7r 



Size of plant horse power 

Cost of plant per horse power. . . . 

Fixed charges at 14 per cent 

Coal per horse-power hour, pounds 

Cost of fuel at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, supplies 

Total 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



700 


800 


900 


1000 


1500 


$76.00 


$69.00 


$64.00 


S60.00 


358.00 


10.65 


9.65 


8.95 


8.40 


8.12 


4.0 


3.5 


3.0 


2.5 


2.0 


22.00 


19.20 


16.50 


13.75 


11. oa 


4.70 


4.15 


3.75 


3.50 


3.25 


0.94 


0.83 


0.75 


0.70 


0.65 


38.29 


33.83 


29.95 


26.35 


23.02 


43.79 


39.73 


34.05 


29.80 


25.77 


41.04 


36.28 


32.00 


28.05 


24.39 


38.29 


33.83 


29.95 


26.35 


23.02 


35.54 


31.48 


27.87 


24.60 


21.64 


32.79 


29.03 


25.80 


22.00 


20.27 


30.04 


27.18 


23.75 


21.20 


18.89 


27.29 


24.23 


21.70 


19.47 


17.52 



2000 

$56.00 

7.85 

1.5 

8.25 

3.00 

0.60 

19.70 

21.75 

20.72 

19.70 

18.67 

17.65 

16.60 

15.57 



FINANCE AND ECONOMICS — COST OF POWER 881 
TABLE 161. 

COST OF POWER. 

Pacific Gas and Electric Company. 

Kilowatt-hours generated by steam 85,707,854 

Kilowatt-hours generated by transmission 7,787,959 

93,495,813 

KUowatt-hours sold 68,797,090 

Kilowatt-hours lost in distribution 24,698,723 

Per cent loss, 26.5. 

TOTAL COSTS. 

Revenue from sales $2,730,248. 00 

Cost of generation $729,315. 00 

Cost of distribution 347,182. 00 

Cost of administration 943,363.00 2,019,860.00 

Net earnings $7 10,388 . 00 

UNIT COSTS, CENTS PER KILOWATT-HOUR. 

Generation: Distribution: 

Labor 0.225 Labor 0.216 

Materials 0.731 Materials 0.098 

Repairs 0.104 Repairs 0.191 

1.060 0.505 

Administration: Summary of Unit Costs: 

Labor 0. 271 Generation 1.060 

Materials 0.082 Distribution 0.505 

Legal Expenses 0.021 Administration 0.576 

Fire Insurance 0.005 Interest 0.006 

Bad Debts 0.026 Depreciation 0.789 

Advertising 0.008 2.936 

Damages to persons . 005 

Rental 0.005 

Taxes 0.153 

0.576 

434. Elements of Power-plant Design. — The real problem which 
confronts the designing engineer is not so much the selection and 
arrangement of apparatus for a given set of conditions as it is to foresee 
the conditions under which the plant is hkely to operate. For this 
reason the plans for the station should be examined and approved by 
an experienced designing engineer, in case expert service is not em- 
ployed at the outset. It is not sufficient to have a mechanically per- 
fect plant, though of course proper installation is of prime importance. 
The choice of fuel, selection of type of prime mover, size of units, 
provision for future expansion, and similar factors bear considerable 
weight upon the economy of operation. Each proposed installation is 
likely to be a problem in itself, and though similar plants may be used 
as patterns, each case should be worked out on its own merits. 

The most important factor in the design of a power station is the 
determination of the probable load curve. This refers not only to the 
average yearly load but also to the maximum daily load which is likely 
to occur, the minimum daily load, temporary peak loads, and probable 



882 



STEAM POWER PLANT ENGINEERING 



02 m 



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FINANCE AND ECONOMICS — COST OF POWER 



883 



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884 



STEAM POWER PLANT ENGINEERING 



future increase. The station load factor and the yearly load factor 
which have such a marked bearing on the cost of operation may be 
closely approximated from the daily load curves. Steam requirements 
for heating and industrial purposes, water supply, and other forms 
of energy requirements should be considered simultaneously with the 

135 
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Curves showing Range 
in Cost of Power 
in 200 Mfg. Plants 

Middle- Western States 


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Size of Plant, Horse -Power 

Fig. 600. 

electrical demands since these factors largely influence the choice of 
prime mover. The curves in Figs. 601 to 603 are taken from the daily 
records of large power stations in Chicago and serve to illustrate the 
great variation in the electrical power demands for different days in 
the year. It is quite evident that at equipment based solely upon the 
average yearly requirements may not be adapted to the best eco- 
nomical operation. 



FINA.NX'E AND ECONOMICS — COST Oi^ POWER 



885 



The load curves for manufacturing plants may be predetermined 
with a fair degree of accuracy since the power demands for various 
purposes may be readily segregated and analyzed, but with public 



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utility concerns and certain classes of isolated stations the problem is 
largely a matter of judgment. Thus, in the case of an industrial 
plant, the power requirements for lighting, manufacturing purposes, 
heating, ventilation, and sanitation may be closeiy approximated since 



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-February-12,-PartIy-cloudy 
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Fig. 602. Typical Daily Load Curves, Tall Office Building, Chicago. 

the size of building, exposure, number of floors, and the number oi 
elevators afford a definite basis for analysis; but with public utility 
concerns the probable load depends largely upon the business acumen 



886 



STEAM POWER PLANT ENGINEERING 



of the management in securing customers, the location of the plant 
and future demands. In the latter case the load curve is based chiefly 
upon the experience of similar plants under comparable conditions of 
operation. 

In any case the greatest care should be exercised in estimating the 
maximum peak load which is likely to occur. High peak loads with 
low daily average necessitate the installation of large machines which 
are idle or operate uneconomically the greater part of the time and 



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Fig. 603. Typical Daily Load Curves, Large Central Station, Chicago. 

result in heavy fixed charges. The financial failure of many electric 
light and power plants is directly traceable to the failure to consider 
the influence of maximum peak loads on the ultimate cost of operation. 
In connection with central-station service every customer represents a 
certain investment, regardless of the amount of power used. Even 
should he consume no power, his account would have to be carried on 
the books and a certain amount of equipment would have to be held 
in readiness to serve him. In order that every customer shall incur 
his share of the expense, the expense of the plant must be apportioned 



FINANCE AND ECONOMICS — COST OF POWER 



887 



between the capacity and output costs. The heavier the peak loads 
the greater will be this charge, and, as is the case with many small 
lighting plants where current is used but three or four hours a day, the 

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Fig. 604. Daily Load Curve showing Influence of Variable Generator Load 
on Steam Economy. 

readiness to serve charge becomes excessive and either the station must 
operate at a loss or the unit cost will appear to be prohibitive. 

The curves in Fig. 604 are taken from recording ammeter and re- 
cording steam meter readings of a 200-kilowatt direct-connected and 
a 45-kilowatt belted generator set installed at the power plant of the 
Armour Institute of Technology and serve to illustrate the influence of 



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Institute of Technology. 

load on economy for very unfavorable conditions. At 8:00 a.m. the 
small unit is started up with initial load of about 150 amperes. As the 
load increases the water rate decreases, as is shown by the curve AB. 
At 9: 00 A.M. the load is beyond the capacity of the small machine and 



888 



STEAM POWER PLANT ENGINEERING 



the large unit is put into service. Tiie increased water rate of the large 
unit over the requirements of the smaller is apparent by the sudden 
rise in the water-rate curve. This is due to the fact that the large unit 
is operating at only 20 per cent of its rating, against full load for the 
small one. The fluctuation of the water rate with the load variation 
is clearly shown. Evidently thQ two units are not of the proper size 
for the particular load conditions illustrated in Fig. 590. During the 
heating months when Hve steam is necessary for "make-up" purposes 



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Fig. 606. Yearly Load Curve showing Influence of Temperature on Coal Con- 
sumption, Combined Heat and Power Plant, Armour Institute of Technology. 

the unfavorable engine load has little effect on the ultimate economy, 
but during the summer months the loss from this cause may be a serious 
one. 

The curves in Figs. 604 to 606 show that during the winter months in 
a combined heat and power plant the fuel requirements may be prac- 
tically uninfluenced by the electrical demands and increase in electrical 
output does not effect an appreciable increase in fuel consumption, but 
the influence of the outside temperature is clearly indicated. 



BIBLIOGRAPHY. — COST OF POWER. 

Accounting Systems for Electric Companies, Power . 38: 697 Nov. 11, 1913 

Air Compressing by Electricity, Cost of. Power. ... 39: 153 Feb. 3, 1914 

Analysis of Plant Costs, Practical Engineer 18: 89 Jan. 1, 1914 

Analysis of Industrial Power Costs, Power 33: 950 June 20, 1911 

A Schedule of Rates Involving the Consumer, 

Electrical World 57: 1562 June 15, 1911 

Boiler Room Economics, Bui. Kan. State Agricul- 
tural College 41: 44 Dec, 1914 

Central Station, Cost of Power in. Power 36: 700 Nov. 12, 1912 

Elec. World 65: 1915 Jan. 30, 1915 

Kansas City Plants, Electrical World 67: 1103 May 13, 1916 



FINANCE AND ECONOMICS — COST OF POWER 889 

Central Station, Massachusetts Plants, Nat. Engineer : 72 Feb., 1916 

Massachusetts Plants, Elec. World 67: 1418 June 17, 1916 

Massachusetts Plants, Elec. World 66: 303 Aug. 7, 1915 

Central Station Rate Making, Power 42: 268 Aug. 24, 1915 

Comparison of Electric Light and Power Rates, 

Power 42: 8 Jan. 6, 1915 

Comparative Cost of Gas and Steam Plants, Power . . 38: 365 Sept. 9, 1913 

Controlling Cost of Electricity, Nat. Engineer : 325 June, 1915 

Cost Data of Power Plant Installation and Oper- 
ation, Eng. Magazine 42: 549 Jan., 1912 

Cost Keeping in the Power Plant, Prac. Engineer. . . 18: 85 Jan. 1, 1914 

Co-Relationship of the Factors Affecting the Cost 

of Power, Eng. Magazine March, 1916 

Cost Accounting in a Modern Hotel, Prac. Engineer 19: 259 Mar. 1, 1915 

Department Store, Power Costs in, Power 39: 122 Jan. 27, 1914 

Detail Plant Costs, Prac. Engineer 19: 735 Aug. 1, 1915 

Distribution Loss Factors, Power 44: 408 Sept. 19, 1916 

Electrical Railway Power Stations, Cost Data, 

Elec. World 68: 865 Oct. 28, 1916 

Electric Rate Making, Nat. Engineer : 164 Mar. 1914 

Equipment, Some Costs of Power Plant, Prac. 

Engineer 18: 1064 Nov. 1, 1914 

Exhaust Steam Heating, Cost of. Power 36: 788 Nov. 26, 1912 

Hotel Buildings, Power Costs in: 

Records at the Blackstone, Power 43: 551 Apr. 18, 1916 

La Salle, Chicago, Power 40: 628 Nov. 3, 1914 

Initial and Operating Costs of Power Plants, Power 44: 304 Aug. 29, 1916 

Isolated Power Plant Records, Prac. Engineer 17: 355 Apr. 1, 1913 

Isolated Stations, Power Costs in: 

Cambridge Y. M. C. A., Power 39: 684 May 12, 1914 

Falk Co., Milwaukee, Power 36: 814 Dec. 3, 1912 

Federal Bldg., Chicago, Power 41 : 610 May 4, 1915 

General, Power 35: 460 Apr. 2, 1912 

Hotel Buildings, Prac. ^ng^ineer 17:834 Sept. 1, 1913 

Kansas City Bldg., Prac. Engineer 17: 1037 Sept. 1, 1913 

Metz Furniture Co., Prac. Engineer 20: 201 Feb. 15, 1916 

Renton Hell Company, Power 38:456 Sept. 30, 1913 

Small Mfg. Plant, Power 41:51 Jan. 12, 1915 

Small Mfg. Plant, Elec. World 66: 1314 Dec. 11, 1915 

N. Y. Hall of Records, Power 43:361 Nov. 14, 1916 

N. Y. Hall of Records, Power 43:597 Apr. 25, 1916 

Paper Mill, Power 44: 190 Aug. 8, 1916 

Why the Isolated Plant Should Win, Power. ... 40: 846 Nov. 15, 1914 

Municipal Plants, Power Costs in: 

Cleveland, Power 41 : 104 Jan. 19, 1915 

Detroit, Power 40: 832 Dec. 15, 1914 

Fort Wayne, Power 45: 546 Apr. 24, 1917 

Kalamazoo, Power 41:218 Feb. 16, 1915 

Office Buildings, Power Costs in, Power 40: 681 Nov. 10, 1914 

Office Buildings, Power Costs in. Power 40: 71 Aug. 18, 1914 

Union Central Building, Power 43:206 Feb. 15, 1916 



890 STEAM POWER PLANT ENGINEERING 

Power Costs, General, Prac. Engineer 18: 99 Jan. 1, 1914 

Power Costs, General, Power 40: 567 Oct. 20, 1914 

Power Costs, General, Prac. Engineer 19: 735 Aug. 1, 1915 

Power Costs, General, Eng. Magazine 48: 278 Nov., 1914 

Public Service Electric Rates, Prac. Engineer 18: 877 Sept. 1, 1914 

Public Service Electric Rates, Prac. Engineer 18: 1109 Nov. 15, 1914 

Records, Power Plant: 

Department Store, Prac. Engineer 21: 275 Mar. 15, 1917 

Detroit Edison, Power 42: 343 Sept. 7, 1915 

General, Power 36: 704 Nov. 12, 1912 

Office Building, Power 43: 15 Jan. 4, 1916 

Packing Plant, Prac. Engineer 17: 355 Apr. 1, 1913 

Reducing of Manufacturing Costs, Prac. Engineer . . 20: 201 Feb. 15, 1916 
Scientific Management in Power Plants, Prac. 

Engineer 19:1 Jan. 1915 

Standardization of Power Plant Operating Costs, 

Jour. A.S.M.E 38: 290 Apr. 1916 

Steam Costs in 6600-hp. Boiler Plant, Power 41: 368 Mar. 16, 1915 



PROBLEMS. 

1. The rated capacity of a steam turbine station is 2000 kw. If the annual out- 
put is 6,380,000 kw-hr., required the yearly load factor. 

2. If the plant in Problem 1 operates 18 hr. per day for 300 days in the year,'re- 
quired the station or curve load factor. 

3. If the plant in Problem 1 cost $65.00 per kw. of rated capacity and the annual 
fixed charges amount to 14 per cent, required the fixed charges per kw-hr. 

4. A plant cost originally $100,000.00. It is proposed to establish a sinking 
fund on a 3 per cent basis. If the weighted life of the plant is assumed to be 20 
years and the junk value of the apparatus at the expiration of this period is estimated 
at 15 per cent of the original cost, how much money must be placed in the reserve 
fund each year. 

5. What will be the accumulated fund in Problem 4 at the end of 15 years? 

6. A steam plant erected 10 years ago at a cost of $250,000.00 is to be appraised 
for rate making. The average weighted life of the equipment is estimated as 25 years. 
What is the accrued depreciation and the present value of the plant on the "straight- 
line" basis. Salvage assumed to be 10 per cent of the original cost. 

7. The average fuel consumption of a 30,000-kw. turbo-generator plant is 2.2 
lb. coal (11,000 B.t.u. per lb.) per kw-hr. for a yearly load factor of 0.42. The cost 
of coal is 2.00 per ton of 2000 lb. and the fuel cost is 45 per cent of the total station 
operating costs. What is the total cost of operation, dollars per year? 

8. A 20,000-kw. turbo-generator uses 14 lb. steam per kw-hr., initial pressure 
215 lb. absolute, superheat 150 deg. fahr., vacuum 27.5 in. referred to a 30-in. ba- 
rometer, feed water 180 deg. fahr. If the average overall boiler and furnace effi- 
ciency is 70 per cent and the calorific value of the coal is 12,500 B.t.u. per lb., re- 
quired the average B.t.u. supplied by the fuel per kw-hr, generated. Determine 
also the average weight of coal used per kw-hr. 

9. During the winter months all of the exhaust steam from a 500-hp. non-condens- 
ing engine is used for heating purposes. Engine uses an average of 60 lb. steam per 
kw-hr., initial pressure 125 lb. abs., back pressure 17 lb abs., initial quality 98 per 
cent, feed water 210 deg. fahr. If the average overall boiler and furnace efficiency 
is 65 per cent and the coal costs $3.00 per ton of 2000 lb. (calorific value 12,000 B.t.u. 
per lb.), what is the actual cost of fuel for power only, cents per kw-hr? 



CHAPTER XIX 



TYPICAL SPECIFICATIONS 



435. Specifications for a Horizontal Tubular Steam Boiler. * — The 

following specifications for one 72-inch horizontal return tubular steam 
boiler, pressure 150 pounds, were prepared by the Hartford Steam 
Boiler Inspection and Insurance Company for the Armour Institute 
of Technology, Chicago: 

This specification is intended to cover the construction of one hori- 
zontal tubular boiler designed to operate at a maximum pressure of 
150 pounds per square inch. Each bidder must submit a proposal 
for doing the work exactly as specified but alternate proposals involv- 
ing slight modifications will also receive consideration provided such 
modifications are fully described. 

The Boiler Contractor shall furnish the various accessories men- 
tioned herein and he shall also provide all the necessary miscellaneous 
iron or steel work as hereinafter enumerated. The Contractor under 
this specification will not be required to construct foundations, brick- 
work or other masonry. 

Drawings. — Drawings prepared by The Hartford Steam Boiler 
Inspection and Insurance Company accompany this specification and 
are made a part hereof; the drawings and specification are intended 
to supplement each other and to be mutually co-operative, and, unless 
otherwise noted, the Boiler Contractor shall follow all details and shall 
furnish all parts and fittings which may be required by the drawings 
and omitted by the specification, or vice versa, just as though required 
by both. The said drawings are identified respectively by Nos. 6260 
and 4890. 

General Data. — The boiler with its fittings shall be constructed and 
furnished in accordance with the following general data and 
dimensions : — 

Diameter measured on inside of largest course 72 inches. 

Number of courses Three. 

Thickness of material : Heads, fs inch. Butt-straps, y^^ inch. Shell- 
plates, ^^ inch. 

Girth seams : Single-riveted lap-joints with rivets spaced 2| inches 
on centers. 

Longitudinal seams Quadruple-riveted butt-joints. 

Diameter of rivets for all seams J inch (jf-inch holes). 

Tubes : Number, 70. Diameter, four inches. Length, 18 feet. 
Thickness, 0.134 inch. 

* Paragraphs pertaining to properties of steel plates, rivets, and tubes have been 
greatly abridged because of space limitation. 

891 



892 STEAM POWER PLANT ENGINEERING 

Braces above tubes : Number on each head, 20. Least diameter, 
1| inches. Diameter of rivet holes for attaching, | inch. Least 
cross-sectional area through sides at each rivet hole on head end, 
0.55 square inch; ditto on shell end 1.10 square inches. 

Through-braces below tubes : Number, 2. Least diameter, two inches. 
Least diameter of upset on front end, 2J inches. Diameter of 
pin. If inches. Least cross-sectional area through center of eye, 
3.83 square inches. 

Size of blow-off pipe 2^ inches. 

Diameter of nozzles : Steam opening 6 inches. 

Safety valve connection 6 inches. 

Size of feed-pipe IJ inches. 

Manholes : One in front head below tubes and one in top of shell. 

Size of grates 72 inches long by 66 inches wide. 

Height from grates to bottom of shell, at front end 40 inches. 

Smoke-Box : Bolted to front head by clip angles. Smoke opening 
60 inches by 14 inches. 

Style of Front Flush. 

Fittings to be furnished with the boiler as follows: — One ten-inch 
steam gauge graduated from to 225 pounds, brass siphon and 
union-cock for gauge, two 2J-inch safety valves with minimum 
lift of 0.08 inch, flanged Y-base for safety valves, three f-inch 
gauge cocks, one combination water-column, one J-inch gauge glass 
14 inches long. 

Method of Support. — The boiler shall be suspended by means of 
U-bolts and steel hangers, from a framework made up of four I-beams 
and four columns. I-beams shall be eight inches deep and shall weigh 
18 pounds per foot; they shall be assembled in pairs by means of tie- 
bolts and separators, spaced near each end and at intervals of not 
more than four feet, in such manner that the adjacent edges will be 
three inches apart. If cast-iron columns are used they shall be round 
with an outside diameter of eight inches and a thickness of J inch, or 
square with a width of eight inches and a thickness of f inch. Six- 
inch rolled steel H-beams, weighing 23.8 pounds per foot, may be used 
for columns but no other form of structural steel column will be ap- 
proved unless it can be shown that the safe load (figured in the usual 
manner with regard to length and radius of gyration) will be equal to 
that which can be allowed on the H-beams specified above. Steel 
columns shall have suitable base-plates and cap-plates riveted on and 
cast-iron columns shall be made with top and bottom flanges of proper 
design. Details of hangers, U-bolts, etc., are shown on the accompany- 
ing drawing. 

Properties of Steel Plates. — (Chemical requirements have been 
omitted.) Complete tests must be made to show that each plate will 
fulfill the above requirements in regard to tensile strength, elastic 
limit, chemical composition, elongation, bending, and homogeneity; 
and any plates failing to meet the said requirements shall be rejected. 
One tension, one cold-bend, and one quench-bend test shall be made 
from each plate as rolled. All details in regard to size and shape of 
specimens, method of making tests, etc., shall be in strict accordance 



TYPICAL SPECIFICATKJNS 893 

with the ''Requirements for Testing Steel," as adopted by The Hart- 
ford Steam Boiler Inspection and Insurance Company. 

All tests and inspections of material may be made at the place of 
manufacture prior to shipment. Certified copies of reports of all 
tests must be approved by a representative of The Hartford Steam 
Boiler Inspection and Insurance Company before any of the material 
covered thereby is used for any portion of the work contemplated by 
this specification. 

Stamping. — (Omitted.) 

Rivets. — (Omitted.) 

Details of Riveting. — Longitudinal seams shall be of the butt-joint 
type with double covering straps and the details shall be as specified 
herein and as shown on the accompanying drawing, except that the 
pitch of rivets in the outer row may be increased or decreased (with 
corresponding changes in the pitch of rivets in the other rows) in cases 
where such changes are desirable in order to secure a proper spacing 
of rivets between girth seams. It must be understood, however, that 
no such change can be made without the consent and approval of the 
inspector having jurisdiction and no such change shall be allowed if 
it will result in a factor of safety lower than 5.00 or if it will produce 
a pitch too great for proper calking. Except for rivet holes in the 
ends of butt-straps, the distance from the center of the rivet to the 
edge of the plate must never be less than one and one-half (1|) times 
the diameter of the rivet hole. The seams must be arranged to come 
well above the fire-line and to break joints in the separate courses. 

Rivet, holes shall either be drilled full size with plates, l)utt-straps 
and heads bolted up in position or else they shall be punched at least 
one-quarter inch {\") less than full size. If the latter method is used, 
plates, straps, and heads shall be assembled and bolted together after 
punching and the rivet holes shall be drilled or reamed in place one- 
sixteenth inch dV") larger than the diameter of the rivets. After 
reaming or drilling, plates and butt-straps shall be disconnected and 
the burrs removed from the edges of all rivet holes. If any holes are 
out of true more than one-sixty-fourth inch {-i^"), they must be brought 
into line with a reamer or drill; evidence that a drift-pin has l^een used 
for this purpose will be sufficient cause for the rejection of the entire 
work. The plates must be rolled to a true circle l)efore drilling and the 
butt-straps and ends of plates forming the longitudinal joints must be 
formed to the proper curvature by pressure, — not by blows. Par- 
ticular care must be used to secure proper fitting where the courses 
telescope together at girth seams. This is a matter of the utmost 
importance and the results obtained will be considered as a criterion 
of the general character of the workmanship throughout. 

Rivets must be of sufficient length to completely fill the rivet holes 
and form heads equal in strength to the bodies of the rivets. Rivets 
shall be machine driven wherever possible, and always with sufficient 
pressure to entirely fill the rivet holes; the authorized inspector of 
The Hartford Steam Boiler Inspection and Insurance Company shall 
have the privilege of cutting out rivets to see if satisfactory results 
have been obtained and all such work of cutting rivets and replacing 



894 STEAM POWER PLANT ENGINEERING 

them shall be done at the expense of the Contractor. Rivets shall be 
allowed to cool and shrink under pressure. 

Calking and Flanging. — All lacking edges shall be beveled to an angle 
of about fifteen degrees (15°) and every portion of such edges shall be 
planed or milled to a depth of not less than one-eighth inch (|"). Bevel- 
shearing will not be acceptable in place of planing or milhng but chip- 
ping will be allowed in special cases provided the workmanship will 
meet with the inspector's approval. All seams must be carefully calked 
with a round-nosed tool. 

Flanging must be performed in such manner that the flange will 
stand accurately at right angles to the face of the sheet and the straight 
portion of the flange must be long enough to allow for making a perfect 
joint with the shell plate. The radius of the bend, on the outside, 
shall be at least equal to four times the thickness of the head. 

Tubes. — (Chemical requirements and method of testing have been 
omitted.) Each tube must be legibly stenciled with the name or brand 
of the manufacturer, the material from which it is made (steel or char- 
coal iron), and the words ''Tested at 1,000 lbs." 

All tests and inspections shall be made at the place of manufacture 
and the Boiler Contractor shall require the tube manufacturer to certify 
that the tubes have been tested and have met the requirements stated 
above. Tubes shall be rejected when inserted in the boiler if they 
fail to stand expanding and beading without showing cracks or flaws, 
or opening at the weld. 

Tube holes may either be drilled full size or punched so as to have 
a diameter at least one-half inch (Y') less than full size and then 
drilled, reamed, or finished full size with a rotating cutter. The full 
size diameter of the hole shall be ^V ii^ch greater than the outside tube 
diameter. Edges of tube holes shall be properly chamfered. 

Tubes shall be set with a Dudgeon expander and all ends shall be 
substantially beaded. 

Staying. — The number, size, arrangement, and general details of 
stays or braces are specified on page — and shown on the drawing. 
No changes shall be made in the number and location of braces without 
the approval of The Hartford Steam Boiler Inspection and Insurance 
Company. All braces shall be made of soUd, weldless mild steel. 

Braces above the tubes shall be of the diagonal crowfoot form and 
none of them shall be less than three feet, six inches (3' 6") long. 
Each brace shall be attached by means of four rivets, two at each end; 
rivets of a larger diameter than specified on page — may be used if 
preferred, but the cross-sectional area through the brace at the sides 
of the rivet holes must be maintained as called for. Braces having 
a rectangular cross-section may be used provided the cross-sectional 
area of each brace is equal to that of each of the round braces specified, 
and provided also that the requirements regarding size of rivets and 
net area through rivet holes are fulfilled. Braces must be carefully 
set to bear uniform tension. 

Through braces shall be used below the tubes, extending from head 
to head. Each brace shall be upset on the rear end to form an eye 
and the eye shall be inserted between the outstanding legs of a pair 
of angle-irons and held in place by a turned bolt passing through holes 



TYPICAL SPECIFICATIONS 895 

drilled in both angles and in the eye. The angles shall be securely 
riveted to the rear head in the manner shown on the drawing, being 
held at a distance of three inches from the head by means of spacers 
made of extra heavy pipe. Spacers must be accurately squared on 
both ends so that they will all be of the same length and will furnish 
a rigid and uniform bearing for the angles. Through braces shall be 
upset and threaded on the front ends and shall pass through the front 
head, being secured with nuts and washers both inside and outside. 
The center Hne of the braces at the front head must not be lower than 
the center line of the manhole. 

Manholes. — Manholes shall be oval or elHptical in shape, not smaller 
than fifteen inches long by eleven inches wide, and shall conform to 
the following requirements: — 

The manhole in the top of the shell shall be placed with its long 
dimension crossways of the boiler. The frame shall be made of pressed 
steel formed to the proper curvature, and it shall be riveted to the 
inside of the shell with two rows of rivets symmetrically spaced. Based 
on the allowance of 44,000 pounds per square inch the size and number 
of the rivets must be such that their total shearing strength will not be 
less than twice the tensile strength of the plate removed, as figured from 
the cross-sectional area in a plane passing through the center of the 
manhole and the axis of the shell; the net cross-sectional area of the 
manhole frame, as cut by such a plane, must not be less than the cross- 
sectional area of the plate removed in the same plane. 

The manhole in the front head shall be formed by flanging the head 
inwardly to a depth of not less than three times the thickness of the 
head all around the opening and a steel band shall be shrunk on, pinned 
in position, and properly machined for the gasket bearing; the band 
will not be required if a recessed manhole plate is used. 

All necessary manhole plates, yokes, bolts, and gaskets shall be fur- 
nished to make the installation complete, the various parts being 
proportioned so as the have a strength equal to that of manhole frames. 
Manhole plates and yokes shall be made of pressed steel. Gasket 
bearings shall be at least one-half inch (i") wide and the thickness of 
gaskets shall not exceed one-quarter inch (i")- 

Nozzles. — Nozzles shall be made of pressed or cast steel and shall 
be of heavy^ and substantial design properly adapted to the pressure 
to be carried. They must be accurately shaped to fit the curvature 
of the shell and must be carefully and securely riveted in place in such 
manner that the face of each flange after erection will lie in a hori- 
zontal plane parallel with the upper surface of the tubes. The flange 
of each nozzle must be properly faced. 

Feed Piping. — Feed piping must be firmly supported in the boiler in 
such manner that no portion of the piping can be in contact with any 
of the tubes or other parts of the boiler. The feed-pipe shall enter the 
boiler through the front head by means of a brass or steel bushing 
placed on the left-hand side of the boiler, three inches (3'0 above the 
top of the upper row of tubes as shown on the drawing. The feed- 
pipe shall extend back from the bushing to approximately three-fifths 
the length of the boiler, crossing over to the center and discharging above 
the tubes. The pipe must not discharge in proximity to any riveted joint. 



896 STEAxM POWER PLANT ENGINEERING 

All external feed-piping will be furnished under separate contract 
but the Boiler Contractor must leave the threads in proper condition 
so that the piping can be readily connected. 

Blow-off Pipe Connection. — A connection for blow-off pipe shall be 
provided on the bottom of the shell near the rear end, as shown on the 
drawing. It shall consist of an extra-heavy pressed steel flange, prop- 
erly tapped for the blow-off pipe and securely riveted to the boiler 
shell. 

Fusible Plug. — A fusible plug shall be placed in the rear head, on the 
vertical diameter, and the center of the plug must not be less than 
two inches (2") above the upper surface of the tubes. The plug must 
project through the sheet not less than one inch (I")- 

Fusible plugs shall be filled with pure tin the least diameter of which 
shall be one -half inch (i"). 

Safety Valves. — Safety valves shall be of the direct spring-loaded 
pop type with seats and discs of nickel or other non-ferrous material. 
Valves must operate without chattering and must be set and adjusted 
to close after blowing down not more than six pounds (6 lb.). Springs 
must not show a permanent set exceeding ^^^ inch ten minutes after 
being released from a cold compression test closing the spring solid; 
no spring shall be used for a pressure in excess of ten per cent (10%) 
above or below that for which it was designed. 

Each safety valve shall have a substantial lifting device with the 
spindle so attached that the valve disc can be lifted from its seat 
through a distance not less than one-tenth of the nominal diameter 
of the valve, when there is no pressure on the boiler. 

The following items shall be plainly stamped or cast upon the body: 

(a) The name or identifying trade-mark of the manufacturer. 

(b) The nominal diameter with the words ''Bevel Seat" or "Flat 

Seat." 

(c) The steam pressure at which the valve is set to blow. 

(d) The lift of the valve disc from its seat, measured immediately 

after the sudden lift due to the pop. 

(e) The weight of steam discharged in pounds per hour at the pres- 

sure for which it is set to blow. 
(/) The letters A.S. M.E. Std. 

Safety valves having a lower lift than that specified on page may 
be used but the diameter must be increased proportionately as directed 
by The Hartford Steam Boiler Inspection and Insurance Company. 

In the absence of any specific directions from the Purchaser, the 
Boiler Contractor shall state in his proposal the make and style of 
valve which he intends to furnish. It is understood that failure to 
do this will give the Purchaser the right to specify the make of valve 
after the contract is awarded and, in such event, the Contractor agrees 
to furnish any make the Purchaser may select. 

Fittings. — The foregoing in regard to choosing the make and style 
of safety-valves shall apply in the same manner and with equal force 
to the make of gauge-cocks, water-column, steam-gauge, etc. 

The combination type of water-column shall be used and openings 
for water and steam connections must be tapped for one-and-one- 



TYPICAL SPECIFICATIONS 897 

quarter-inch (Ij") pipes. Brass pipe shall be provided for the water 
connection and the piping shall be made up with plugged fittings to 
facilitate cleaning. 

The Boiler Contractor shall properly drill and tap all hok^s required 
for the installation of the various fittings, including also a one-quarter- 
inch (j'O pipe with valve for the connection of test gage. The sizes 
of steam-gauge, gauge-cocks, and gauge-glass are specified on page 0. 

All nozzles, flanges, fittings, etc., furnished under this specification 
must correspond in diameter, drilling, and other details with the 
''American Standard" for the stipulated pressure. 

Front. — The front shall be constructed of sectional plate steel or of 
cast iron and the Contractor must state in his proposal which form 
he intends to furnish. If made of steel, the plates must not be less 
than three-eighths inch (§") thick (except for moldings, etc.) and 
they must be straight and smooth with all edges machined and properly 
fitted to make good joints. Heavy cast-iron door-frames with planed 
surfaces shall be securely bolted to the plates and the frcmt shall be 
further reinforced against warping by means of channel irons or other 
suitable braces placed on the back. 

If made of cast-iron, the front must be of heavy and substantial 
design and all castings must be smooth, true, and free from cracks, 
blow-holes, or other defects. 

The usual fire-doors, ash-pit doors, and doors for giving access to 
the tubes shall be provided as shown on the accompanying drawings. 
All doors must be of heavy design and all contact surfaces must be 
carefully machined so that the doors will fit closely. Each flue door 
must be provided with a suitable fastening at top and bottom, de- 
signed to clamp the door tightly in the closed position and prevent 
warping. All doors shall ])e furnished complete with handles, catches, 
hinge-bolts, etc., and fire-doors shall have liner plates. 

The Boiler Contractor shall furnish all necessary anchor bolts for 
holding the front in position and shall see that the holes for the same 
are properly located in the steel plates or castings. Anchor bolts 
shall have a diameter of at least seven-eighths inch (}'') and shall be 
threaded and provided with nuts. 

All parts must be carefully made so that the front will present a 
neat appearance after erection. Open joints, loosely-fitting hinges or 
other indications of careless workmanship will be sufficient cause for 
rejection and the Purchaser shall have the option of making any 
necessary modifications and deducting the cost thereof from the 
contract price or of requiring the Contractor to furnish new parts which 
will be satisfactory. 

Grates. — The Boiler Contractor shall figure on furnishing stationary 
grates of suitable design and shall base his proposal thereon. If re- 
quested by the Purchaser, he shall submit an alternate proposal for 
furnishing, shaking, rocking, or dumping grates of a type wiiich the 
Purchaser will specify. 

Miscellaneous Iron Work. — Arch-bars for rear connection shall be 
made as shown on the accompanying drawings or in accordance with 
some detail which will meet with the approval of The Hartford Steam 
Boiler Inspection and Insurance Company. The Company will not 



898 STEAM POWER PLANT ENGINEERING 

approve any arch-bar the metal of which is exposed to the action of 
the flames and hot gases. 

The rear connection door must fit closely and the frame must be 
provided with means for anchoring into the brickwork. The door 
must not be smaller than sixteen inches by twenty-four inches (16" 

X 24'0. 

The Boiler Contractor shall furnish all necessary bearer-bars for 
grates, all buckstays, tie-rods, lintels for clean-out doors, bolts, etc., 
and any other iron-work, not specifically mentioned herein, which may 
be needed to complete the installation in the brick setting. Buck- 
stays must be made of pressed steel or its equivalent; cast-iron will 
not be accepted. 

Tests. — The Boiler Contractor shall at all times afford all facilities 
to The Hartford Steam Boiler Inspection and Insurance Company, 
and its authorized representatives, for the test and inspection of all 
materials and workmanship entering into the work covered by this 
specification. 

Hydrostatic tests shall be made in the presence of the authorized 
inspector of The Hartford Steam Boiler Inspection and Insurance 
Company and in a manner which will meet with the approval of the 
said inspector. The pressure for such tests shall not exceed one and 
one-half (1|) times the maximum working pressure as hereinbefore 
stated. 

Local or State Laws. — All details of construction and installation 
shall be made in strict accordance with any local or State ordinances 
which may apply and nothing in this specification shall be interpreted 
as an infringement of such rules or ordinances. If any discrepancy 
should arise, the Contractor shall immediately report it to The Hartford 
Steam Boiler Inspection and Insurance Company for settlement. 

436. Specifications for Steam, Exliaust, Water, and Condenser Piping 
for an Electric Power Station.* — The work referred to in this contract 

shall be conducted under the general supervision of 

(referred to as the Engineers), who shall interpret the Specifications 
and the Drawings that may accompany the Specifications, and shall 
arbitrate any controversies between the parties hereto, that may arise 
under this contract, their decision to be final and binding upon both 
of the contracting parties. 

The Contractor shall comply with all laws, statutes, ordinances, 
acts, and regulations of the town or city, the state and the government 
in which the work is to be performed, and shall pay all fees for permits 
and inspections required thereby. 

The Contractor shall, at an early date, communicate with other 
contractors employed by the Purchaser, and shall work in harmony 
with them, any differences of opinion between contractors being arbi- 
trated by the Engineers or their representative. 

The Contractor shall begin work as soon as possible, and complete 
same, free of all liens and charges, on or before the time mentioned 
herein. If, in the opinion of the Engineers, the Contractor fails to 
prosecute the work with the necessary means and diligence to insure 

* From the files of a prominent Chicago engineering firm. 



TYPICAL SPECIFICATIONS 899 

its completion within the time Umit, then the Engineers shall notify 
the Contractor by written notice to that effect, and the Purchaser 
may order the Contractor to employ more men, machinery, and tools 
to be put upon the work, specifying the additional force required, and 
if the Contractor fails to comply with such written demand within 
six (6) days from the date thereof, or within such time as the Engineers 
in writing prescribe, then the Purchaser may employ necessary means 
to complete the work within the time required, and such additional 
cost caused by either the employment of additional men, machinery, 
or otherwise, shall be deducted from any funds due, or that may be- 
come due the Contractor on account of this contract. The Con- 
tra,ctor shall remove any particular workman or workmen from the 
work, if in the judgment of the Engineers it will be for the best interest 
of the work. 

The Engineers shall have the right to make any changes in the 
Drawings or Specifications that they deem desirable. Should any 
additional labor or material be involved in such changes, the Con- 
tractor shall be paid for supplying same; on the other hand, should 
such changes reduce the amount of labor or material from that origi- 
nally specified, the Contractor shall sustain an equivalent reduction 
in the contract amount and the Engineers shall be the arbiters in 
determining rates of increase or reduction. No claim shall be allowed 
for extra labor or material above the contract amount, unless same 
shall have been ordered in writing, with remuneration stipulated, by 
the Engineers. Acceptance by the Contractor of final payment on 
the contract price shall constitute a waiver of all claims against the 
Purchaser. 

All material and workmanship furnished under this contract must 
be of the best quality in every particular and the Contractor must 
remedy any defects which develop during the first year of actual 
service, due to faulty material or workmanship, free of expense to the 
Purchaser. The Purchaser, the Engineers, or their representative may 
inspect any machinery, material or work to. be furnished under this 
contract and may reject any which is defective or unsuitable for the 
uses and purposes intended, or not in accordance with the intent of 
this contract, and may order the Contractor to remedy or replace 
same; or the Purchaser may, if necessary, renjedy or replace same at 
the expense of the Contractor. 

Until accepted in its entirety by the Purchaser, all work shall be 
done at the Contractor's risk, and if any loss or damage should occur 
to the work from fire or any other cause, the Contractor shall promptly 
repair or replace such loss or damage free of all expense to the Pur- 
chaser. The Contractor shall be responsible for any loss or damage 
to material, tools or other articles used or held for use in or about the 
work. 

The work shall be carried on to completion without damage to any 
work or property of the Purchaser or of others, and without inter- 
fering with the operation of their machinery or apparatus. 

The Contractor shall furnish all false work, tools and appliances 
that may be required to accomplish the work and shall remove all 
debris after erection. 



900 STEAM POWER PLANT ENGINEERING 

The Contractor must be responsible for the safety of the work until 
finished and accepted by the Purchaser and must maintain all lights, 
guards, and temporaiy passages necessary for that purpose. In case 
cf any accident causing injury to person or property, the Contractor 
shall obtain acquittance from or pay the injured person (whether such 
person be an employee, a fellow-contractor, an employee of a fellow- 
contractor, or otherwise) the amount of damages to which he or she 
may be legally entitled on account of any act or omission of the Con- 
tractor or of any agent or employee of the Contractor, during the 
performance of the work referred to herein, and shall provide adequate 
insurance to protect the Purchaser from all claims arising therefrom. 
The Contractor shall, further, insure the compensation provided for in 
any workman's compensation act which may affect the work, to all 
its employees or their beneficiaries, and the Contractor shall carry 
insurance in a company satisfactory to the Purchaser, insuring said 
compensation to its employees or their beneficiaries. The Contractor 
shall notify his insurance company and cause the name of the Pur- 
chaser to be incorporated in the compensation pohcy, the policy or a 
copy thereof to be deposited with the Purchaser upon request. The 
Contractor must save the Purchaser harmless from all claims for 
damages set up by reason of any such injury and from all expenses 
resulting therefrom. 

No certificates given or payments made shall be considered as con- 
clusive evidence of the performance of this contract, either wholly or 
in part, nor shall any certificate of payment be construed as acceptance 
of defective work or improper materials. The Contractor agrees to 
furnish the Purchaser or the Engineers, if requested, at any time 
during the progress of the work, a statement showing the Contractor's 
total outstanding indebtedness for material and labor in connection 
with the work covered by this contract, such statement to be certified 
to by a notary public. Before final payment is made the Contractor 
shall satisfy the Purchaser by affidavits or otherwise, that there are 
no outstanding Kens for labor or materials against the Purchaser's 
premises by reason of any work done or materials furnished under 
this contract. 

If, during the progress of the work, the Contractor should allow any 
indebtedness to accrue for labor or material to sub-contractors or 
others, and should fail to pay and discharge same within five (5) days 
after demand made by any person furnishing such labor or material, 
then the Purchaser may withhold any money due the Contractor until 
such indebtedness is paid, or apply same toward the discharge thereof. 

All royalties for patents, or charges for the use or infringement 
thereof, that may be involved in the construction or use of any ma- 
chinery or appliance referred to herein, shall be included in the contract 
price, and the Contractor must satisfy all demands of this nature that 
may be made against the Purchaser at any time. 

This contract shall not be assigned nor shall any part of the work 
be sub-let by the Contractor without the written consent of the Engi- 
neers being first obtained, but such approval shall not relieve the Con- 
tractor from full responsibility for the work included in this contract 
and for the due performance of all the terms and conditions of this 



TYPICAL SPECIFICATIONS 901 

contract; and in no case shall such ai)proval be granted until such 
Contractor has furnished the Purchaser witli satisfactory evidence 
that the Sub-contractor is carrying ample workmen's compensation 
insurance to the same extent and in the same manner as is herein 
provided to be furnished by the Contractor. 



General Data. 

The work herein referred to comprises the furnishing of all material 
and labor for the complete installation of Piping Systems for two (2) 
kw. units to be installed in the Power Station being erected by 

Each of the two (2) units is comprised of the following machinery: 
(List of machinery omitted.) 

All of the above machinery will be installed on the foundations by 
their respective contractors, and this Contractor shall make all piping 
coimection to same unless otherwise mentioned. 

Drawings. (These have been omitted.) 

This contractor shall take such measurement at the building and 
allow for such make-up pieces as shall be necessary to make his work 
come true, as the Purchasee and its Engineers cannot be responsible 
for the exact accuracy of the dimensions given on Drawings. 

The Drawings and Specifications must be taken together and any 
work called for in the one or indicated in the other, or such work as 
can be reasonably taken as belonging to the Piping Connections and 
necessary to complete the system, is to be included. 



Live Steam Piping. 

Connections from Boilers. — Each of the eight (8) boilers will be pro- 
vided with two (2) 8-inch steam outlets to which this Contractor shall 
connect an 8-inch angle automatic stop and check valve with 7-inch 
outlet. From these valves Contractor shall provide 7-inch ])oiler 
leads connecting to the steam mains with gate valve at the mains, 
all arranged as indicated on Drawings, Nos. — and — . 

Connections to Turbines. — Contractor shall provide a cast-steel 
manifold at rear of each of the two boilers on each unit on both sides 
of boiler room and connect to these manifolds the two 7-inch leads 
from the four boilers on each unit. From manifold at rear of boilers 
on north side of boiler room on each unit a 14-inch connection shall 
be run across the basement of firing room and connected together 
with 14-inch lead from manifold at rear of boilers on south side of 
boiler room of each unit into an 17-inch pipe, which shall ])e connected 
to the turbines. A 14-inch hydraulically operated valve shall be 
provided on each 14-inch line where they connect together into the 
17-inch turbine lead; a gate valve shall be provided on turbine lead. 

Connections shall be provided complete with cast-steel manifolds, 
valves, drip pockets, pipe lengths and bends, all of sizes and arranged 
as indicated on Drawings, Nos. — and — . 



902 STEAM POWER PLANT ENGINEERING 

Steam Loops. — Contractor shall provide the 12-inch steam loops be- 
tween the steam leads to turbines complete with pipe bends and a 
hydraulically operated gate valve on each end of loop. Hydraulically 
operated gate valves shall also be provided for connecting the future 
loop, all as indicated on Drawings, Nos. — and — . 

Steam to Auxiliaries. — This Contractor shall install a 4-inch auxil- 
iary steam header along division wall between turbine and boiler rooms, 
with connections to manifolds at rear of boilers on south side of boiler 
room with gate valve at each manifold, all arranged as indicated on 
Drawings, Nos. — and — . From the auxiliary header connec- 
tions shall be made to one service pump in condenser well, three 
feed pumps in boiler room, exciter in turbine room, two auxiliary oil 
pumps on turbines and to tempering coils on air washers, as shown 
on Drawings. The steam connection to each of the pumps must be 
provided with angle or globe throttle valve at pump. A gate valve 
must be provided on each connection near header, as indicated on 
Drawings. Each of the three (3) turbine-driven feed pumps will be 
provided with a 3-inch pressure governor by Pump Contractor, which 
this Contractor shall install in the steam hne. The steam-driven 
service pump will be provided with a 2-inch pressure governor by 
Pump Contractor, which this Contractor shall install, providing a 
by-pass with three valves around same, one of which is to be the 
throttle valve, the other two gate valves. This Contractor shall also 
provide a 3-inch steam connection to the exciter, providing a globe 
valve at turbine and gate valve at header. 

On the steam connections to the oil pumps and air washers this 
Contractor must provide a 1-inch extra heavy pressure-reducing valve 
with by-pass around same for each unit. These shall reduce from 
250 pounds to 100 pounds, and a second reducing valve shall be pro- 
vided on connections to air washers reducing from 100 pounds to 10 
pounds. 

Steam from Turbines to Heaters. — The Contractor shall furnish and 
install the 5-inch steam connections from outlet on intermediate stage 
of each turbine to the auxihary exhaust hne connecting to feed-water 
heaters with automatic stop and check valve, regulating valve oper- 
ated by thermostat in feed-water heater, set so as to heat water to 
about 120 deg. fahr., pressure-reducing valve and gate valve at header, 
as shown on Drawings. The exhaust from steam-driven auxiUaries 
will go to the heaters, and it is the intention to take necessary addi- 
tional steam from second stage of turbine to heat the feed water to 
required temperature. 

Steam Connections to Soot Ejectors. — Contractor shall provide a IJ- 
inch steam header lengthwise on each side of boiler room, with con- 
nections to cast-steel manifolds in main steam connections with gate 
valve at north side of boiler room and to auxiliary steam header with 
valve on south side of boiler room. From these IJ-inch headers a 
1-inch connection with globe valve having extended stem shall be 
run to the ejectors in basement, for each of the two divisions of each 
of the eight economizers, all arranged as indicated on Drawings, Nos. 
— , — and — . 



TYPICAL SPECIFICATIONS 903 

Steam Ejectors on Condenser Discharge Pipes. — Contractor shall pro- 
vide a 4-inch ejector on top of each of the two (2) 54-inch condenser 
discharge pipes. These shall be of Schutte & Koerting or other make 
that Engineers may approve. To each of these ejectors Contractor 
shall provide a 1-inch steam connection with valve on both ends of 
line; also run a 4-inch discharge connection to 6-inch bilge pump 
discharge line with gate and check valve on each Hne. 

Supports for Live Steam Piping. — The main supporting beams upon 
which the manifolds and fittings are supported will be provided by 
contractor for building steel, but this Contractor shall furnish the 
steel brackets framing to the main members above mentioned; also 
all roller and anchor bearings, complete with base castings, rollers, 
straps, spring, etc., all as indicated and detailed on Drawings. He 
shall provide the steel frames for supporting the 14-inch steam load 
across the boiler room basement. He shall also provide the bearings 
for supporting the pipes on those supports. This Contractor shall 
also provide the main anchor bearings for the 17-inch steam loads to 
turbines; also the roller bearings and brackets for the 17-inch steam 
load to Unit No. 2. 

The steel brackets for supporting the auxiliary steam header will 
be provided by Contractor for Building Steel, but this Contractor 
shall provide the roller and anchor bearings on these brackets, all as 
indicated on the Drawings. 

Contractor shall also provide such additional hangers, braces and 
supports for the steam piping as may be necessary to properly support 
the steam piping, and keep same free from vibration. These must in 
all cases be of steel or iron, and made subject to the approval of the 
Engineers. 

Steam Drips and Drains. — The main steam headers shall be drained 
to the 10-inch drip pockets in boiler room basement. This Contractor 
shall provide and install a l^-inch steam trap for each unit for drain- 
ing the drip pocket and must connect up same with a l^-inch pipe. 
The discharge from the trap shall be connected to the feed-water 
heater. Connections at trap shall be arranged with by-pass with 
three valves, so trap can be cut out of service. 

Each of the 7-inch gate valves on steam leads from boilers shall have 
a boss tapped for J-inch drain above seat, which this Contractor shall 
connect into a IJ-inch hne for each unit and connect same with stop 
and check valve to the feed-water heater, also to the clear water 
reservoir; IJ-inch lines to be cross connected with valves. Contractor 
shall provide a boss tapped for J-inch drain on the 12-inch hydrauli- 
cally operated gate valves on steam loop, also on the two 14-inch 
valves on lead from manifolds at rear of boilers for each unit, and 
connect same with a 1 J-inch pipe to their respective steam traps, 
providing by-pass with valves as indicated diagrammatically on draw- 
ings. The 12-inch gate valve for future steam loop shall also have 
boss tapped for f-inch drain and connected to the 1 J-inch drain line. 
A globe valve shall be provided on each drain connection. Contractor 
shall also tap the blind flange on tee in steam connection to condenser 
well and provide a f-inch drain connection with trap and discharge 



904 STEAM POWER PLANT ENGINEERING 

connection to the feed-water heater. A by-pass connection with 
three valves shall be provided at trap. A J-inch drain shall also be 
provided from lowest point of steam connection in condenser well to 
drain sump. 

Contractor shall rmi a |-inch drain with valve from the steam 
casing of the three auxihary turbines driving the boiler-feed pumps 
and the turbine driving the exciter and connect them into a 1-inch 
Hne and run to the hot water reservoir. Drain from casing of service 
pump turbine to be run to drain sump in condenser well with a valve 
at turbine. 

Contractor shall also provide such other drip and drain connections 
as may be necessary to properly drain the entire system of steam 
connections, these to be connected as may be directed by the Engineers. 

Blow-off Connections. 

Boiler Blow-off Connections. — Each of the eight boilers will be pro- 
vided with six (6) 2i-inch blow-off fittings on mud drums, which this 
Contractor shall connect up to a special fitting on each side of each 
boiler and from which 2J-inch connections shall be made to the blow- 
off header under each row of boilers. Eight (8) 2i-inch blow-off 
valves shall be provided on the blow-off connection from each of the 
eight boilers, all arranged as indicated on Drawings. 

Contractor shall also provide the 4-inch blow-off header under each 
row of boilers and run 4-inch connections from same to the steel blow- 
off tank in boiler-room basement. This tank will be furnished and 
installed by Contractor for steel tanks, but this Contractor shall pro- 
vide the overflow and drain connections to discharge well and vent 
connections to atmosphere, all of sizes and arranged as indicated on 
the Drawings. 

Superheater Blow-off Connections. — This Contractor shall furnish 
and install the superheater blow-off connections from each of the eight 
boilers to the blow-off header in basement, as indicated on Drawings. 
Each boiler will be provided with two (2) 2-inch elbows and two (2) 
2-inch valves, one on each end of each drum and two elbows and two 
valves on superheater, which this Contractor must connect to the 
headers. Six (6) 2-inch valves must be provided for these connections 
on each boiler, all arranged as indicated on Drawings. 

Blow-off from Economizers. — Each of the eight (8) economizers will 
be pro^dded with eight (8) 2i-inch blow-off outlets, provided with 
angle valves. This Contractor shall connect these together to a 4-inch 
header, providing a 2J-inch valve on each of the two divisions on each 
of the eight economizers. Headers shall be run along just below 
economizer floor, and 4-inch connection shah be run to hot water 
reservoir and 4-inch to discharge line from blow-off tank. A globe 
valve with extended stem shall be provided on each of these connections. 
A check valve shall also be provided where connection is made to dis- 
charge from blow-off tank. On the economizer side of these globe 
valves tee shall be tapped for J-inch pipe and connection run to pet 
cock above boiler-room floor, which shall drain into a funnel connected 
to discharge weU. 



TYPICAL SPECIFICATIONS 905 

Exhaust Connections. 

Exhaust Connections from Turbines. — This Contractor shall furnish 
and install the 42-inch free air exhaust connections from each of the 
two (2) turbines, as indicated on Drawing No. — , made up of 
cast-iron pipe and fittings and riveted steel pipe with forged steel 
riveted flanges, as made by the American Spiral Pipe Works. The 
steel pipe shall be close riveted and thoroughly calked so as to be air 
and water tight. Copper expansion joint shall be provided between 
main turbine exhaust and relief valve on each unit. The vertical 
risers shall be of J-inch plate and shall terminate above roof, with 
hoods over same, as per detail on Drawings. Horizontal pipe between 
relief valve and base elbow shall be of y^-inch steel plate. There is 
to be no longitudinal seam on bottom of this pipe. The exhaust relic:" 
valves in these lines shall be as hereinafter specified under ''Material 
and Workmanship." 

Exhaust Connections from Auxiliaries. — This Contractor shall connect 
up the exhaust outlet on the three (3) turbine-driven feed pumps, 
auxiliary oil pumps, ser\'ice pump and exciter together, and make 
connection to each of the two feed-water heaters, with gate valve at 
each pump, each heater and sectionalizing valve between heaters, all 
of sizes and arranged as indicated on Drawings. A 10-inch riser to 
atmosphere with combination back pressure and relief valve near 

heater and exhaust head above roof shall be provided on 

connections to each of the two heaters. Exhaust heads shall be of 
No. 16 galvanized iron and of most improved type. Each heater will 
also be provided with a 4-inch relief outlet, which this Contractor shaU 
connect up with a back pressure valve to the 10-inch rehef pipe to 
atmosphere on each unit, all arranged as indicated on Drawings. 

Heating Syste7n for Switch House, Operating Room and Offices. — 
Contractor shall furnish and install for heating switch house, operating 
room, and offices, a complete two-pipe heating system, with overhead 
supply system and drain in basement. The switch house heating 
system shall have a total direct radiation of approximately 1912 square 
feet, divided into 17 radiators. The operating room, offices, bedrooms, 
stair hall, etc., at end of turbine room shall have a total radiation of 
approximately 3188 square feet, divided into 55 radiators, all of sizes 
and arranged as may be directed by the Engineers. A layout drawing 
showing size of radiators and sizes of branch connections will be pro- 
vided later. All radiators to be " " 

two-column radiators, or other make that the Engineers may approve. 
All radiators to have top steam connections. 

Steam for this system shall be taken from the auxiliary exhaust 
header in boiler room, with a 6-inch connection running up the stair 
hall to the bus chamber under switch house, with gate valve and 3-inch 
safety valve set at 5 pounds pressure in boiler room. A low-pressure 
header shall be run across the bus chamber and up to the overhead 
header in switch house, which shall be run along the south wall and 
connected to the radiators in switch house. An overhead line shall 
also be run around three sides of the office space over switchboard 
room with drop connections to the radiators on the different floors. 



906 STEAM POWER PLANT ENGINEERING 

Drains from the radiators shall all be brought together and connected 
to a direct-connected, geared, motor-driven vacuum pump as made by 
the American Steam Pump Co. and of ample capacity for the service 
and to maintain a vacuum of 5 inches at the outlet of radiators. Motor 
to be similar to those hereafter specified and must be complete with 
starting equipment switches, fuses, etc. All wiring between motor 
and equipment to be provided. 

Discharge from pump shall be connected to the feed-water heater 
by means of a float-controlled vent, as made by Company. 

A J-inch syphon trap shall be provided on outlet of each radiator, 

as made by , and a standard radiator valve provided on inlet of 

each radiator. All piping to be rigidly suspended in approved manner. 

Safety Valve Vent Pipe. — This Contractor shall furnish and install 
the safety valve vent pipes on each of the eight (8) boilers, as shown 

on Drawings, Nos. . The Discharge openings of the six (6) 

4j-inch safety valves on drum of each boiler shall be connected together 
as indicated, and a 12-inch riser run through roof and terminating 
in a 12-inch tee. He shall also furnish and install the safety valve 
vent pipes from the discharge openings on each of the two (2) 4-inch 
superheater safety valves on each of the eight (8) boilers. The out- 
lets of two valves shall be combined into a 6-inch pipe and run through 
roof terminating in a 6-inch tee. A J-inch drain pipe shall be provided 
on elbows at each safety valve, connecting into a f-inch pipe from each 
boiler, which shall be run to ash pit. 

Exhaust Drips. — This Contractor shall install a 2i-inch drip pipe 
from the 42-inch free exhaust from each turbine, providing a deep 
U-trap and discharging into hot water reservoir under boiler room 
basement floor. 

The Turbine Contractor will connect up the drains from the carbon 
packing rings into a 3-inch pipe on each of the two (2) turbines. This 
Contractor shall connect each of these pipes to the hot water reservoir. 
Gate valves on vertical connections from auxiharies shall be tapped 
above seats for J-inch bleeders, which shall be connected together into 
a 1-inch line and run to hot water reservoir. Drain from gate valve 
on service pump shall be run to drain sump in condenser well. 

Support for Exhaust Piping. — Relief valves on turbine exhaust lines 
shall be provided with bases, which will be supported from floor under 
valves, and the vertical risers will be carried on the base elbows, but 
this Contractor shall provide and set angle iron braces for vertical 
risers, as per detail. 

This Contractor shall provide all necessary anchors, hangers, and 
braces for properly supporting the auxiUary exhaust lines, as may be 
required by the Engineers. 

Water Piping. 

Circulating Water Connections. — Purchaser will provide and install 
the suction connection from intake crib to the suction inlet on each of 
the two circulating pumps. 

Condenser Contractor will provide the discharge connection from 
circulating pump to condenser on each unit. 



TYPICAL SPECIFICATIONS 907 

Purchaser will furnish and install the condenser discharge piping 
outside of consender well, including gate valves, elbows, and vertical 
pipe length in discharge well, but this Contractor shall provide the 
special fitting, pipe lengths, and expansion joints on condenser dis- 
charge connections inside of condenser well. One of the pipe lengths 
on discharge connection from Unit No. 1 in the condenser well will 
be provided on ground by Purchaser, but this Contractor shall install 
same, providing gaskets and bolts for making up joints, all arranged 

and of sizes as indicated on Drawing . Contractor shall also 

provide the 6-inch tail pipes from 54-inch gate valves in discharge well. 

Hot-well Pump Connections. — Contractor shall connect up the two 
hot-well pump discharge outlets on each unit to the inlet on primary 
heater in upper section of condenser, providing check and gate valve 
at each pump. From outlet of primary heater, connection shall be 
run to inlet on top of heater of each unit. The primary heater is also 
to be by-passed with necessary valves, all of sizes and arranged as 

indicated on drawings, Nos. . Connections to heaters shall 

be cross connected with valves as indicated on Drawings. 

Feed Pump Suction Connections. — Contractor shall furnish and in- 
stall the suction connections to the two (2) feed pumps on each unit 
with connections from heater, filtered water header and unfiltered 
water system with valve on each connection, all of sizes and arranged 

as indicated on Drawings, Nos. . Suction connections from 

heaters shall be cross connected with valve as indicated. 

Boiler-Feed Piping. — This Contractor shall furnish and install dis- 
charge connections from the feed pumps to the feed headers and from 
feed headers to economizers and boilers, all arranged as shown on 
Drawings. There are to be two separate feed-water systems for each 
unit with independent connections from pumps to boilers, as shown. 
The auxiliary feed header is to be run in the boiler room at rear end 
between boilers and in basement across firing room to boiler on north 
side of room, with connections from same to boilers. The main feeder 
header shall be suspended from the economizer floor framing over 
boilers with connections to each of the eight (8) economizers and from 
economizers to the boilers. Connections between the economizer 
divisions will be provided by Economizer Contractor. 

Each boiler will have two (2) feed inlet connections and Boiler Con- 
tractor will provide a 4-inch automatic stop and check valve on each 
of these outlets, to which this Contractor shall connect. 

Each economizer will be provided with a 4-inch inlet at bottom and 
a 4-inch outlet at top, which this Contractor shall connect up. 

From the 7-inch auxiliary feed headers, this Contractor shall run a 
4-inch connection up the front of boilers, with a 4-inch connection to 
the inlet at each end of drum, providing a gate valve at header con- 
nection and a globe and check valve in horizontal run at front of boiler. 

From 7-inch main feed headers. Contractor shall make a 4-inch 
connection to each economizer with two gate valves on each connection. 
He shall also make a 4-inch connection from outlet of each economizer 
to the feed line connecting to each of the boilers, providing a gate and 
check valve at economizer outlet and an angle globe valve with ex- 
tended stem all arranged as indicated on Drawings. 



908 STEAM POWER PLANT ENGINEERING 

Contractor shall provide two air chambers on each of the two main 
feed headers, and one air chamber on each of the two auxihary headers, 
with gate valve on headers and with compressed air connections with 
extra-heavy stop and check valves. 

Contractor shall provide a 6-inch cross connection between the two 
(2) 7-inch main feed lines and auxiliary feed lines, with gate valve on 
each connection, as indicated. Connections at pumps shall be ar- 
ranged with special two-way check valves and gate valve, all of sizes 
and arranged as indicated on Drawings. 

Water Connections to Hydraulically Operated Valves. — This Con- 
tractor shall provide and connect up a four-way cock for the hydrauli- 
cally operated valve on the steam lead to turbine; the two 14-inch 
valves on steam lead from boilers; the 12-inch valve on steam loop 
on each unit and the 12-inch valve for future steam loop. The four 
4-way cocks on each unit are to be located in a box set in the division 
wall between boiler and turbine rooms, all as indicated on Drawings. 
Boxes shall also be provided by this Contractor. Water supply for 
the four-way cocks is to be taken from both the feed headers, with 
gate and check valves arranged as indicated on Drawing. Drain 
connections with troughs and drain pipes connected to hot water well 
are to be provided as indicated. 

The following items included in the complete specifications have 
been omitted: 

High-pressure Boiler Washing System. 

Service Water Piping. 

Make-up Water Connections. 

Water Drains. 

Miscellaneous Drains and Vents. 

Oil Connection to Turbines. 

Pipe and Fittings for Oihng Systems. 

Compressed-air System. 

Air Washer Circulating Pump Suction. 

Floor and Wall Thimbles. 

Hose. 

Thermometers and Gauges. 

Material and Workmanship. 

General Instructions. — All material and workmanship supplied under 
these Specifications shall be the best of their respective kinds. 

All material shall be such as specified herein and free from defects 
or flaws of any kind, and subject to such tests and requirements as 
may be herein described or as may be necessary to prove the effective- 
ness of the material or workmanship. All labor is to be performed by 
men skilled in their particular line of work, and to the full satisfaction 
of the Supervising Engineers or their representatives. The Speci- 
fications contemplate the very best quality of material and the most 
mechanical character of workmanship. 

All of the work shall be erected, ready for practical use, to the satis- 
faction of the Engineers, and all bolts, gaskets, and necessary adjunccs 
shall be furnished by this Contractor. 



TYPICAL SPECIFICATIONS 909 

This Contractor shall satisfy himself as to the accuracy of the 
Drawings, and must take such measurements and allow for such 
make-up lengths or pieces as may be necessaiy to make his work come 
accurately together. The piping must be erected so as to preserve 
accurate alignment and no iron gaskets or fillers will be allowed be- 
tween flanges. 

Where the work of this Contractor connects to that of another, the 
connections shall be made by this Contractor, and he must see that 
all flanges for connection to the other work are properly drilled to fit 
the latter, irrespective of drilling dimensions on the Drawings or herein 
given. 

The work contemplated herein shall be carried on so as to harmonize 
and not interfere with the work of other contractors or with the opera- 
tion of the Station or any of the machinery that may be contained 
therein. Where connections are made to the old work, they shall be 
done at such time as shall meet the approval of the Chief Engineer 
of the Station. The work shall be installed as expeditiously as pos- 
sible and subject to the general direction of the authorized Engineers. 

The following items pertaining to material and construction details 
are included in the complete specifications ])ut have l^een omitted 
from this copy. 

Steel Pipe. Traps. 

Welded Flanges. Flanged Joints. 

Threaded Flanges and Unions. Cast-iron Pipe. 

Fittings. Supports and Hangers. 

Valves. Testing. 

Hydraulically Operated Valves. Pipe Covering. 

Relief Valves. Painting. 
Special Valves and Appliances. 

437. Government Specification and Proposal for Supplying Coal. 

U. S. Treasury Department. 

United States 

, 190.. 

PROPOSAL. 

1 Sealed proposals will be received at this office until 2 o'clock p. m., 

2 , 190. . , for supplying coal to the United States 

3 building at 

4 as follows : 

5 

6 

7 

8 The quantity of coal stated above is based upon the previous annual 

9 consumption, and proposals must be made upon the basis of a deliver}^ of 

10 10 per cent more or less than this amount, subject to the actual require- 

11 ments of the service. 

12 Proposals must be made on this form, and include all expenses incident 

13 to the delivery and stowage of the coal, which must be delivered in such 

14 quantities, and at such times within the fiscal year ending June 30, 190 , 

15 as may be required. 



910 STEAM POWER PLANT ENGINEERING 

16 Proposals must be accompanied by a deposit (certified check, when 

17 practicable, in favor of ) 

18 amounting to 10 per cent of the aggregate amount of the bid submitted, as 

19 a guaranty that it is bona fide. Deposits will be returned to unsuccessful 

20 bidders iimnediately after award has been made, but the deposit of the 

21 successful bidder will be retained until after the coal shall have been de- 

22 livered, and final settlement made therefor, as security for the faithful 

23 performance of the terms of the contract, with the understanding that the 

24 whole or a part thereof may be used to liquidate the value of any deficiencies 

25 in quality or delivery that may arise under the terms of the contract. 

26 When the amount of the contract exceeds $10,000, a bond may be exe- 

27 cuted in the sum of 25 per cent of the contract amount, and in this case, the 

28 deposit or certified check submitted with the proposal will be returned after 

29 approval of the bond. 

30 The bids will be opened in the presence of the bidders, their representa- 

31 tives, or such of them as may attend, at the time and place above specified. 

32 In determining the award of the contract, consideration will be given to 

33 the quality of the coal offered by the bidder, as well as the price per ton, 

34 and should it appear to be to the best interests of the Government to 

35 award the contract for supplying coal at a price higher than that named in 

36 lower bid or bids received, the award will be so made. 

37 The right to reject any or all bids and to waive defects is expressly 

38 reserved by the Government. 

DESCRIPTION OF COAL DESIRED.* 

39 Bids are desired on coal described as follows: 

40 

41 

42 

43 

44 

45 

46 

47 

48 

49 

50 Coals containing more than the following percentages, based upon dry 

51 coal, will not be considered: 

52 Ash per cent. 

53 Volatile matter per cent. 

54 Sulphur per cent. 

55 t Dust and fine coal as delivered at point of consumption per cent. 

DELIVERY. 

56 The coal shall be delivered in such quantities and at such times as the 

57 Government may direct. 

58 In this connection, it may be stated that all the available storage capacity 

59 of the coal bunkers will be placed at the disposal of the contractor to 

60 facilitate delivery of coal under favorable conditions. 

61 After verbal or written notice has been given to deliver coal under this 

62 contract, a further notice may be served in writing upon the contractor to 

* Note, — This information will be given by the Government as may be deter- 
mined by boiler and furnace equipment, operating conditions, and the local market, 
t Note. — All coal which will pass through a i-inch round-hole screen. 



TYPICAL SPECIFICATIONS 911 

63 make delivery of the coal so ordered within twenty-four hours after receipt 

64 of said second notice. 

65 Should the contractor, for any reason, fail to comply with the second 

66 request the Government will be at liberty to buy coal in the open market, 

67 and to charge against the contractor any excess in price of coal so purchased 

68 over the contract price. 

SAMPLING. 

69 Samples of the coal delivered will be taken by a representative of the 

70 Government. 

71 In all cases where it is practicable, the coal will be sampled at the time 

72 it is being delivered to the building. In case of small deliveries, it may be 

73 necessary to take these samples from the yards or bins. The sample 

74 taken will in no case be less than the total of one hundred (100) pounds, to 

75 be selected proportionally from the lumps and fine coal in order that it 

76 will in every respect truly represent the quality of coal under considera- 

77 tion. 

78 In order to minimize the loss in the original moisture content the gross 

79 sample will be pulverized as rapidly as possible until none of the fragments 

80 exceed | inch in diameter. The fine coal will then be mixed thoroughly 

81 and divided into four equal parts. Opposite quarters will be thrown out, 

82 and the remaining portions thoroughly mixed and again quartered, throw- 

83 ing out opposite quarters as before. This process w^ill be continued as 

84 rapidly as possible until the final sample is reduced to such amount that 

85 all of the final sample thus obtained will be contained in the shipping can or 

86 jar and sealed air-tight. 

87 The sample will then be forwarded to the Chief Clerk of the Treasury 

88 Department, care of the storekeeper. 

89 If desired by the coal contractor, permission will be given to him, or his 

90 representative, to be present and witness the quartering and preparation of 

91 the final sample to be forwarded to the Government laboratories. 

92 Immediately on receipt of the sample, it will be analyzed and tested by 

93 the Government, following the method adopted by the American Chemical 

94 Society, and using a bomb calorimeter. A copy of the result will be mailed 

95 to the contractor upon the completion thereof. 



CAUSES FOR REJECTION. 

96 A contract entered into under the terms of this specification shall not 

97 be binding if, as the result of a practical service test of reasonable duration, 

98 the coal fails to give satisfactory results due to excessive clinkering, or to 

99 a prohibitive amount of smoke. 

100 It is understood that the coal delivered during the year will be of the 

101 same character as that specified by the contractor. It should, therefore, 

102 be supplied, as nearly as possible, from the same mine or group of mines. 

103 Coal containing percentages of volatile matter, sulphur, and dust higher 

104 than the hmits indicated on line 54, and coal containing a percentage of 

105 ash in excess of the maximum limits indicated in the following table, will 

106 be subject to rejection. 

107 In the case of coal which has been delivered and used for" trial, or which 

108 has been consumed or remains on the premises at the time of the deter- 

109 mination of its quality, payment will be made therefor at a reduced price 

110 computed under the terms of this specification. 

111 Occasional deliveries containing ash up to the percentage indicated in 

112 the column of ''Maximum limits for ash," on page 912, may be accepted. 



912 



STEAM POWER PLANT ENGINEERING 



113 Frequent or continued failure to maintain the standard established by 

114 the contractor, however, will be considered sufficient cause for cancellation 

115 of the contract. 

116 Payment will be made on the basis of the price named in the proposal 

117 for the coal specified therein, corrected for variations in heating value and 

118 ash, as shown by analysis, above and below the standard established by 

119 contractor in this proposal. For example, if the coal contains two (2) 

120 per cent, more or less, British thermal units than the established standard, 

121 the price will be increased or decreased two (2) per cent accordingly. 

122 The price will also be further corrected for the percentages of ash. For 

123 all coal which by analysis contains less ash than that established in this 

124 proposal a premium of 1 cent per ton for each whole per cent less ash will 

125 be paid. An increase in the ash content of two (2) per cent over the 

126 standard established by contractor will be tolerated without exactmg a 

127 penalty for the excess of ash. When such excess exceeds two (2) per cent 

128 above the standard established, deductions will be made from price paid 

129 per ton in accordance with following table : 



* PRICE AND PAYMENT. 





No 

deduc- 
tion for 
limits 
below. 


Cents per ton to be deducted. 


Maxi- 


Ash as estab- 
lished in 
proposal. 


2 


4 


7 


12 


18 


25 


35 


mum 

limits 

for 




Pel 


"centages 


of ash in 


dry coal. 






ash. 


Per cent. 
5 


7 
8 
9 

10 
11 
12 

13 
14 
15 
16 
17 
18 
19 
20 


7- 8 

8- 9 
9-10 

10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 


8- 9 
9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 


9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 


10-11 
11-12 
12-13 
13-14 
14-15 
15-16 


11-12 
12-13 
13-14 


12-13 
13-14 
14^15 
15-16 
10-17 
17-18 
18-19 
19-20 
20-21 
21-22 


13-14 
14-15 
15-16 
16-17 

17-18 


12 


6 


13 


7 


14 


8 

9 


14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


14 
15 


10 


16 


11 

12 


16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


16 

17 


13 


18 


14 


19 


15 


19-20 
20-21 
21-22 
22-23 


19 


16 






20 


17 






21 


18 








22 















* Note. — The economic value of a fuel is affected by the actual amount of com- 
bustible matter it contains, as determined by its heating value shown in British 
thermal units per pound of fuel, and also by other factors, among which is its ash 
content. The ash content not only lowers the heating value and decreases the 
capacity of the furnace, but also materially increases the cost of handling the coal, 
the labor of firing, and the cost of the removal of ashes, etc. 

Proposals to receive consideration must be submitted upon this form and contain 
all of the information requested. 

..'...'/.'/...'.\'.'.'.'.y.'.'.'.'.'/.'/.'.'.'.^m 

The undersigned hereby agree to furnish to the U. S 

building at , the coal described, in tons 

of 2240 pounds each and in quantity, 10 'per cent more or less than that stated 
on page 912, as may be required during the fiscal year ending June 30, 190 , 



TYPICAL SPECIFICATIONS 913 

in strict accordance with this specification; the coal to be delivered in such 
quantities and at such times as the Government may direct. 

Price per ton (2240 pounds) $ 

Conunercial name of the coal 

Name of the mine or mines 

Location of the mine or mines 

Name or other designation of the coal bed or vein 

Size (indicate information which will apply) — 
Unsized Lump Run of mine 

{Round 1 .. 
Square J '^P^^^^gs- 
Bar screen. 
Data to establish a basis for pai/ment: 

British thermal units in coal as delivered 

Ash in dry coal (Method of American Chemical Society) per cent. 

It is important that the above information docs not establish a higher standard than 
can be actually maintained under the terms of the contract; and in this connection it should 
be noted that the small samples taken from the mine are invariably of higher quality than 
the coal actually delivered therefrom. It is evident, therefore, that it will be to the best 
interests of the contractor to furnish a correct description with average values of the coal 
offered, as a failure to maintain the standard established by contractor will result in de- 
ductions from the contract price, and may cause a cancellation of the contract, while de- 
liveries of a coal of higher grade than quoted will be paid for at an increased price. 

Signature : 

Address : 

Name of corporation, 

Name of president, 

Name of secretary, 

Under what law (State) corporation is organized: 



CHAPTER XX 

TYPICAL CENTRAL STATIONS 

438. The advancements that are being made in the design of large 
central stations and central station machinery are so rapid that it is 
futile to apply the term "modern" to any installation with the assur- 
ance that the plant thus designated will be representative of current 
practice for even a brief period of time. To-day it is possible to install 
in a given space approximately five times the capacity that could be 
installed a few years ago, with the cost per unit capacity only about 
one fifth and with very little increase in cost per square foot of floor 
space occupied. That the limit has not been reached is evidenced by 
the fact that boiler pressures of 350 lb. per sq. in. are to be employed 
in several plants in course of construction and even higher pressures 
have been considered for future designs. A few years ago boiler 
capacities during peak loads of 250 per cent rating were considered 
exceptional ; to-day 400 per cent and even 500 per cent rating has been 
obtained with high overall efficiency. Improvement has not been 
limited to boilers and prime movers, but has been extended to all 
parts of the equipment. 

The Essex Station of the Public Service Electric Co. of New Jersey, 
the Niagara River Station of the Buffalo General Electric Co., Buffalo, 
N. Y., the Northwest Station of the Commonwealth Edison Co., 
Chicago, may be considered the latest (1917) achievements in power 
plant design; every detail necessary to promote efficient operation 
and continuity of service has been incorporated. 

Essex Power Station. — The plant is built on the unit system in 
what may be considered four separate structures: Switchhouse, tur- 
bine room, boiler house, coal bunkers and coal bridge. The four 
buildings occupy a total frontage of 401 ft. 

The top of the coal tower is 215 ft. above high water and it has a 
lift of 156 ft. The tower with the bunkers is on the east side of the 
boiler room and is equipped with a 600-hp. hoisting engine of the two- 
drum type, and has a capacity of 240 tons per hr. when using a 2-ton 
clamshell bucket. The hoisting speed is 1300 ft. per min. When the 
bucket is dropping, it is driving the motor as an induction generator 
and pumping back into the fine. The hoisting engine is driven by a 
600-hp. induction motor. 

914 



TYPICAL CENTRAL STATIONS 



915 







H 



916 



STEAM POWER PLANT ENGINEERING 




':^^^<':^^m 



TYPICAL CENTRAL STATIONS 



917 



i^Blow-off 




.'V-- ^^I'-r--^: r-'. 



Fig. 609. Essex Station. — Front Elevation, Boiler Equipment. 



918 STEAM POWER PLANT ENGINEERING 

Coal, after being hoisted to the tower, goes into a hopper from which 
it passes through a feeder into the crusher and is then distributed by 
a belt conveyor to the bunkers. The conveyors are driven by induc- 
tion motors and are automatically cut out by push-buttons placed in 
convenient locations along the conveyor runway. Provisions are also 
made for unloading coal directly from cars or barges to the storage 
yard and for reclaiming the coal from the storage yard. All the coal- 
handhng machinery is driven by induction motors. 

The bunkers have a capacity of 2000 tons and are built of rein- 
forced concrete supported on the steel structure of the building. From 
the outside bunkers the coal is brought by means of 15-ton motor- 
driven weighing larries, one in each firing aisle, and distributed to the 
stoker hoppers. Each hopper will hold seven tons of coal which is 
weighed automatically as it is distributed from the larry. 

The boiler room contains eight 1373-hp. cross-drum marine- type 
Babcock & Wilcox water-tube boilers working under a steam pressure 
of 225 pounds. They contain 672 tubes 4 inches by 18 feet, arranged 
42 tubes wide by 16 tubes high, giving a heating surface of 13,723 
square feet. The boilers are guaranteed to evaporate 41,200 pounds 
of water from and at 212 deg. fahr. per hour and will give 300 per 
cent rating with clean heating surfaces. There are also two rows of 
circulating tubes which connect the upper ends of the front headers 
to the steam drum, as indicated in Fig. 609. Each boiler is equipped 
with six 4.5-inch Crosby safety valves arranged in three pairs so as to 
blow into a common header, which is piped through the roof. They 
are also equipped with 2 steam-flow meters, 2 steam gauges, 2 water 
columns, 2 feed-water regulators and 2 feed-water inlets. 

The firing is done with 16-retort underfeed Sanford Riley stokers. 
The drive equipment for each firing aisle consists of four 12-hp. four- 
speed motors, two driving the mainshaft through a jack-shaft and two 
driving through Reeves conical variable-speed transmissions, giving 
a mainshaft speed of 32 to 290 r.p.m. This is equivalent to a coal 
feed of from 1600 to 15,000 lb. per hr. for each boiler. The furnace 
has an active grate area of 200 square feet. This gives a ratio of grate 
area to heating surface of 1 : 63.5. The tubes are 7 feet 10 inches above 
the grate at the curtain wall and 9 feet at the back wall. Each boiler 
is equipped for forced, natural or induced draft, or all three may be 
used at the same time. Forced draft for each boiler is obtained by a 
60,000 cubic feet multivane fan, driven by a 150-hp. motor, which can 
maintain a 6-inch water pressure under the grates. The air supply 
to the furnaces is controlled by a Mason regulator. The air pressure 
under the grates acts upon a flexible diaphragm, which through gears 



TYPICAL CENTRAL STATIONS 



919 



o€^ 



shifts the screws on the Reeves drive, consequently regulating the rate 
at which coal is fed to the furnace. 

Induced draft is obtained by a 100,000-cubic-feet-per-minute multi- 
vane fan located at the economizer outlet and driven by a 100-hp. 
motor. The fan gives a 2-inch suction in the uptake of the boilers. 
After the gases have passed from the boiler, they may go directly to 
the stack, or, by closing dampers 
in the breechings, can be made 
to pass through the economizer 
and then to the stack; by clos- 
ing a second damper, the gases 
will pass through the induced- 
draft fans before going to the 
stack. This makes it possible to 
operate the boilers under the 
most economical conditions at 
all times. 

The economizers contain 480 
four-inch tubes 12 feet long, giv- 
ing a heating surface of 7750 
square feet, which is 56 per cent 
of the boiler-heating surface. 
They are built to stand a work- 
ing pressure of 300 pounds per 
square inch. 

The stacks are two in number, 
250 feet high above the grates 
and 16 feet inside diameter. 
They are built of steel plates 
and are of the self-supporting 
type. They have a 4-inch stack- 
brick lining with from one to 
two inches of grout between the 
brick and the shell. 




Fig. 610. 



Essex Station. — Steam 
Piping. 

Drips and drain lines from all sources, as well as the leaders from the 
roof, discharge into concrete storage tanks beneath the boiler-room 
floor, and the water so collected is used for boiler make-up. The 
balance of the make-up water is taken from the city mains. The con- 
densate from each main unit passes through a metering tank provided 
with a V-notch recording meter having a capacity of 4,000,000 pounds 
per hour and drops by gravity to a 77,000-pound storage tank. The 
make-up water passes through another V-notch recorder, and also 



920 



STEAM POWER PLANT ENGINEERING 



goes to the storage tank with the condensate, whence it passes to two 
10,000-hp. open feed-water metering heaters and then to the boiler- 
feed pumps at a temperature of 164 deg. fahr. and is pumped through 
the economizers into the boilers at a temperature of 244 deg. fahr. 
The feed water to each boiler is controlled by two Copes feed-water 
regulators, which maintain the water level in the boilers constant. 
A regulator maintains constant pressure difference between feed pres- 
sure and steam pressure to turbines on feed pumps. The layout of 
the feed- water-pi ping system and economizers is given in Fig. 611. 
The system is self-explanatory; it will be seen that the layout is very 
simple for what may be accomplished with it and the last word in 
flexibility. There are three boiler-feed pumps. Each pump is 3-stage 



Boiler No . 8 Boiler No . 6 Boiler No . 4 Boiler No . 2 

*1 ^Feedwater— "XLL- ^^^'eedwater /J2 1_ 23^Feedwaver /l_ilj "-^Feedwaler ^^*»i 

" f Regulator 7 ^^ ? Regulator^ f ^ I ' ^ Regulator^ T T Regulator^ f ] 

4.%'i' V •=} 4" '4' 4" "o 4'_^ 1; 4" <=^ 4" _ 'i' 4" 

/ — '' / " h k — A — - — ]—^ — - — j - i X " '\ ''\ — .")« 

IS S ^> S § o^ g / g g "To Testing Line ^ ^ \ ^ Z^ 'ID S / £ N ^ 



"I Feed '^ 
H Pump 
J 5o.3 ' 



" lo Testing Line 

'if»o Feed Pump Xo.l5 

Feed Pump Xo.2 ^_, 



ir ^>, a <^ 1 ; V o = ■ reea rump .-^o.i NV=> ^ =• ^ t 

4" r 4" '^ 4"' 4" "^ 4' } , '*' '' 4" ^ 4" 

J Feedwater ta-Urf Feedwater L^ I '* .f Feedwater f , i\f Feedwater 

a ^X Regulalor^_^ 3;Li|. ^^ Regulator-^,^^^ ] ;^. ^ Regulator— ~^^ 3l4 *^''^'B"'°'°''~ ' 

if V M y Pipe 6 and over, of Semi-Steel Ij V \J 

I l .. 4V Steel ~ i, J l| 

Boiler No. 7 Boilei No. 5 " ° "°° Boiler No. 3 I Boiler No. 1 



^ , f , ^ j 

£j Feedwater U 
^J^-'^' Regulator- ~^jt3' 



Fig. 61L Essex Station. — Feed-water Piping. 

double suction having a capacity of 1000 gallons per minute against 
a total dynamic head of 700 feet and is driven by a 250-hp. Westing- 
house turbine running 2200 r.p.m. 

The boilers blow down through 6 blowdown pipes equipped with 
a Babcock & Wilcox and Everlasting valve in series into a tank, thence 
through a V-notch meter into the sewer. 

Very elaborate arrangements have been made in the piping system 
whereby any turbine or boiler may be tested while in regular service. 

All the soot from the economizers, stacks, etc., is taken out by steam 
ejectors. The ashes drop from the grates into hoppers, where they 
fall into side-dump pivot cars and are hauled out into the yard by a 
5-ton electric locomotive and are used for filhng-in purposes. Pro- 
visions have been made so that when the ashes are no longer required 
for filling purposes, they will be raised by a skip hoist to a bunker in 
the coal tower and from there discharged into barges. 

A switchboard is located in front of each boiler, from which are 



TYPICAL CENTRAL STATIONS 921 

controlled the stoker drive, forced- and induced-draft fans. Indi- 
cating and recording meters are also mounted on these panels for 
draft, temperature, motor current, and stoker speed. 

All boiler stop valves, sectionalizing valves, and turbine stop valves 
are operated by 125-volt, direct-current motors and are controlled 
from a switchboard, in the boiler-room engineer's office on the main 
floor of the boiler room. The turbine stop valves can also be operated 
from a remote-control station on the main floor of the turbine room. 

The steam is taken from a double-ended superheater through 
Edwards stop and check valves into two 8-inch pipes, one at each 
end of the boiler, down through the boiler-room floor to 12-inch to 
18-inch double headers as shown in Fig. 610. These headers are 
cross-connected at each connection from the boilers. A 16-inch steam 
line goes to each turbine and an 8-inch to the auxiliaries. Each end 
of the superheater is furnished with a 4.5-inch safety valve. 

The steam headers are anchored at the center so that one-half of 
the expansion is in each direction from that point. The headers are 
also anchored in the turbine-room basement before they connect to the 
turbine inlet through expansion bends and risers. There are no expan- 
sion joints in the headers; they are installed under a tension between 
anchorages, which causes an elongation equal to about one half of the 
expansion of the section normal temperature to that of the steam. 
Therefore, when the headers are at the temperature of the surrounding 
air they are in tension, and when at the temperature of the steam they 
are in about the same amount of compression. By this scheme it has 
been possible to do away with expansion joints in the headers, and so 
far they have worked very satisfactorily. The headers are carried on 
sling rods with stirrups resting on springs to allow for come and go. 

The turbine room contains two 25, 000-kilo volt-ampere General 
Electric main units, only one being operated at a time. Three boilers 
are required to supply steam for one unit, which gives a ratio of boiler 
to engine horsepower of 1:8. The main turbines are 12-stage, tandem- 
compound, with 8 stages for the high pressure and 4 for the low 
pressure, and exhaust into a surface condenser of the two-pass type 
containing 6434 one-inch tubes 19 feet active length. This gives 1.28 
square feet of cooling surface per kilowatt and provides for the con« 
densation of 7.5 pounds of steam per square foot of cooling surface per 
hour. An average vacuum of 28.73 inch is maintained with 70-degree 
circulating water. The condensers are of the radial-flow type and are 
rigidly connected to the turbines; the expansion and contraction is 
taken care of by supporting the condensers on springs. 

The circulating water is supplied by two 24,000-gallon centrifugal 



922 STEAM POWER PLANT ENGINEERING 

pumps for each main unit, one motor-driven and the other turbine- 
driven. In the winter the turbine-driven pump usually has capacity 
enough to maintain the vacuum, making it unnecessary to run th^ 
motor-driven pump except during the summer months. The electri- 
cally-driven pump is so arranged that when the temperature of the 
discharge water rises above a certain value a thermostat closes, auto- 
matically starting the motor, and puts the second pump into service. 

The vacuum pumps are of the Westinghouse Le Blanc type and are 
motor-driven. A turbine-driven hot-well pump takes care of the con- 
densate, which is pumped back into a tank whence it passes through 
a V-notch meter to the feed-water heater. The exhaust of all auxil- 
iaries goes to the open feed-water heaters. 

The circulating water is taken from the river through three intake 
tunnels 9 feet 4 inches by 8 feet, equipped with motor-driven revolv- 
ing screens. The discharge tunnels are two in number, 12 feet 6 inches 
by 9 feet 4 inches, and rest on top of the intake tunnels. 

The main generator units are 25,000-kilovolt-ampere (continuous 
rating), 60-cycle, three-phase, 13, 200- volt machines running 1800 r.p.m. 
They are equipped with 100-kilowatt, 250-volt, direct-connected ex- 
citers. These are the only generators of this speed and capacity that 
have direct-connected exciters. 

The air for ventilating the generators is taken from outside the 
buildings and is washed by water sprays, one on each generator, 
directed into the incoming air in several directions. This not only 
cleans the air, but also cools and humidifies it, the drop in temperature 
in some cases being as much as 15 deg. fahr. The air is forced through 
the ventilating ducts of the generator by fans on the rotor. The 
heated air from the generator is carried back to the forced-draft fan- 
room in the boiler house and supplies part of the air for the furnaces, 
thus recovering some of the losses in the generator. 

The exciter system is designed to secure maximum reliability to- 
gether with independent excitation for each generator, and consists 
of a regular, emergency, and spare. Regular excitation is supplied by 
a 250- volt shunt generator directly connected to each alternator shaft. 
In case of trouble on the regular exciter, a low-voltage relay instantly 
closes the emergency-exciter circuit, which consists of a 900-ampere 
(30-minute rating) storage battery equipped with four 14-point end- 
cell switches, after which the direct-connected exciter is cut out auto- 
matically by a reverse current relay. The spare exciter, which is a 
75-kilowatt, motor-driven shunt generator, may then be started by 
the operator from the main switchboard sind cut in parallel with the 
battery on the field of the generator and the battery cut out. 



TYPICAL CENTRAL STATIONS 



923 



The power for the station is supphed from a 3000-kilo volt-ampere, 
13,200- to 440-volt, 60-cycle, water-cooled, oil-insulated transformer. 
The transformer is equipped with water-flow indicators and ther- 
mometers, which operate an alarm in case the transformer has no 
cooling water or becomes overheated. 

All the motors used throughout the plant except those which operate 
valves are 440-volt machines. A 440-volt system was selected on 
account of the greater safety to the attendants over a 2300-volt system 
and also on account of the great saving in cable and bus capacity 
compared with 220-volt. Where variable speed is required of the 
alternating-current motors, it is obtained either by changing the number 
of poles in the stator winding or by rotor resistance. 



TABLE 163. 

ESSEX STATION - GENERAL DATA. 

Coal Handling. 



6 

;5 


Equipment. 


Kind. 


Size. 


Operating Condition. 




Hoisting engine 


Two-drum 


24-in. drums, 200 r.p.m. 


180 tons per hour 




H. E. motor 


Induction 


410hp.,200r.p.m. 


Direct connected to 




Traversing engine 


Single-drum 


30-in. drum 


Gear-driven motor 




T. E. motor 


Induction 


75 hp. 


Constant speed 




Feeder 


Apron 


5 ft. wide, 16 ft. long 


Gear-driven motor 




Feeder motor 


Induction 


7.5 hp. 


Constant speed 




Crushers 


2-roll 


36 in. by 36 in. rolls 


Gear-driven motor 




Crusher motors 


Induction 


35 hp. 


Constant speed 




Conveyor 


Belt 


30 in. by 161 ft. long 


300 tons per hour 




Conveyor 


Belt 


30 in. bv 120 ft. long 


200 tons per hour 




Conveyor motor 


Induction 


20 hp. 


Constant speed 




Conveyor motor 


Induction 


10 hp. 


Constant speed 




Coal bucket 
Automatic skip hoist 


Clamshell 
Balanced 


U ton 

2 ton buckets — 100 tons per hr. 






Full automatic 




Skip hoist motor 


Induction 


25 hp. 


Constant speed 




Elevator 


Electric freight 


7^ ton 


50 ft. per minute 




Elevator motor 


Induction 


40 hp. 


Constant speed 



Turbine House. 



2 Generators 



Exciters 
Exciter 
Exciter motor 
Turbines 

Condensers 
Circulating pumps 
C. P.turbines 

C. P. motors 
Air pumps 
Air pump motors 
Condensate pumps 
C. P. turbines 

Crane 

Air washers 

Oil pumps 

Oil pump motors 

Oil filter 



General electric 

Direct-connected 
Shunt wound 
Induction 
Curtis — horizontal 

Two pass surface 
Horizontal-centrifugal 
Horizontally split 

Induction 
Le Blanc 
Induction 
Centrifugal 
Horizontally split 

Electric traveling 
Spray system 

Centrifugal 
Induction 



25,000 kv.a. 

100 kw., 250 v., compound wound 
150 kw., 250 V. 
220hp. — 1175 r.p.m. 
25,000 kw. — 12 stage 

32,000 sq. ft. — 255,000 lb. per hr. 
24,000 gal. per min. 43 ft. max. hd 
350 hp. 

350 hp. 

37.5 cu. ft. free air per min. 

100 hp. 

500 gal. per minute 

20 hp. 

100 ton — 94 ft. span 
60,000 cu. ft. air per minute 

3-in. — 150 gal. per minute 
7.5 hp. 



132,000 v., 3 hp.,60cy., 

1800 r.p.m. 
1800 r.p.m. 
Reserve equipment 
Reserve equipment 
190 lb. steam, 150 deg. 

superheat 

Turbine driven 

190 lb. steam, 150 deg. 

superheat 
Constant speed 
Motor driven 
Constant speed 
Turbine driven 
190 lb. steam, 150 deg. 

superheat 
2-50 ton, 1-10 ton, hooks 
Motor-driven pumps — 

10 hp. 
Motor driven 
Constant speed 
For new & make-up oil 

only 



924 



STEAM POWER PLANT ENGINEERING 



TABLE 16S. — Continued. 
Boiler House. 



Boilers 

Superheaters 
Soot cleaners 
Stokers 

Reeves transmissions 
Stoker motors 
Forced-draft blowers 
F. D. B. Motors 
Economizers 
Induced-draft fans 

I. D. F. motors 
Heaters 
Metering tank 
Metering tank 
Boiler-feed pumps 
B. F. P. turbines 

Fire pump 
Fire pump motor 
Air compressor 
A. C. motor 
Service pumps 
S. P. turbine 

Air compressor 

A. C. motor 

Coal larries 

Stacks 

Locomotives 

Meters 

Charging equipment 



B. & W. cross drum 

Flash type 
Steam blow 
Underfeed (Riley) 



Induction 

Turbo vane (Sturtevant) 

Induction 

C. I. tube (Sturtevant) 

Multivane (Sturtevant) 

Induction 
Cochrane metering 
Blow-off 
Feed-water 
3-stage centrifugal 
Horizontal 

2-stage centrifugal 
Induction 
Straight line 
Induction 
Single stage 
Horizontal 

Straight line 

Induction 

Weighing 

Steel-brick lined 

Electric storage battery 

Steam flow 

Motor generator 



1373 hp. on 10 lb. per sq. ft. basis 
1711 sq. ft. heating surface 



retort — 15,000 lb. coal per hr. 
Class F. — No. 6i 
12 hp. 

000 cu. ft. min. at 7 in. st. pres. 
150 hp. 

7750 sq. ft. 40 by 12 bv 12 ft. 
106,000 cu. ft. gases 400 deg. fahr, 

100 hp. 

10,000 hp., 500,000 lb. per hour 

200,000 lb. per hour 

1 ,000,000 lb. per hour 

1000 gal. per minute 

250 hp. 

1000 gal. per minute 

100 hp. 

225 cu. ft. per minute 

500 gal. per minute 
22 hp. 

100 cu. ft. per minute 

15 hp. 

15 ton 

15 ft. 4 in. by 250 ft. 



15 kw. 



225 lb. press., 150 deg. 

superheat 
150 deg. superheat 
Live steam 
Continuous dumping 
Variable 
Four speed 
2 speed, motor driven 
2 speed 

1 economizer per boiler 
Constant speed motor 

driven 
Constant speed 



700 ft. total head 

190 lb. steam, 150 deg. 

superheat 
110 lb. total head 
Constant speed 
100 lb. pressure 
Constant speed 
110 lb. total head 
190 lb. gauge, 150 deg. 

superheat 
100 lb. pressure 
Constant speed 
Motor driven 



battery 



For elec. locomotives 



TABLE 164. 

BUFFALO GENERAL ELECTRIC CO. — GENERAL DATA. 

Boiler Room 



and 



Wilcox, cross-drum water tube 
5 



Type of boilers Babcock 

Number now installed 

Anticipated station load, immediate, kw 

Heating surface, each, sq. ft 

Superheater surface, each boiler, sq. ft 

Grate surface per boiler, total, sq. ft 

Heating surface per sq. ft. grate surface, sq. ft. . . 

Heating surface per sq. ft. superheater surface, sq. ft 

Superheater surface per sq. ft. grate surface, sq. ft 

Heating surface per kw. (95,000 kw., five 11,400 sq. ft. boilers) 

Working pressure, lb. per sq. in 

Superheat, deg. fahr 

Total temperature steam, deg. fahr ( 

Mud drum material Forged 

Stokers, Riley underfeed, 2 per boiler; retorts per boiler 

Maximum capacity each retort, lb. coal per hr 

Capacity each retort at approx. 400 per cent, rating, lb. per hr 

Boiler rating on peaks, per cent, 350; between peaks, per cent 

Rating on peaks, anticipated maximum, lb. water per hr. from and 

at 212 deg. fahr 

Water evaporated per sq. ft. heating surface at approx. 400 per cent 

rating, lb. per hr 

Capacity each stoker on peaks, kw 5,000 to 10,000 

Coal burned per sq. ft. grate at 250 per cent rating, lb. per hr., 26; 

at approx. 400 per cent rating, lb. per hr 50. 25 



40,000 

11,400 

3,815 

418 

27.27 

3 

9.1 

0.6 

275 

275 

689.4 

steel 

30 

1,000 

700 

100 to 250 

160,300 

14.4 



TYPICAL CENTRAL STATIONS 



925 



TABLE 164. — Continued. 

Economizers, green, type H, maximum pressure, lb. per sq. in 

Economizer material 

Economizer heating surface per boiler, sq. ft 

Economizer heating surface per sq, ft. boiler-heating surface, sq. ft. 
Economizer surface per boiler horsepower (34.5 lb. water per hr.) 

rated capacity, sq. ft 

Economizer — Guarantees for each unit economizer : 



400 

Cast iron 

9,435 

1.208 

0.228 



Feed Water, 
Lb. per Hr. 


Temperature Leav- 
ing Economizer 
when Entering at 
ISO Deg. Fahr. 


Gas Temperature 

Entering, Deg. 

Fahr. 


Gas Temperature 

Leaving, Deg. 

Fahr. 


Gas Temperature 

Difference, De?. 

Fahr. 


53,000 

86,000 

103,000 

120,000 


263 
281 

289 
288 


535 
633 
670 
705 


294 

366 
396 
443 


241 

267 
274 
262 



200 



200 



Coal Bituminous run of mine 

Coal bunker, type Non-suspended, steel frame, concrete lined 

Coal conveyors: Two bucket conveyors, capacity each per hr., tons 
Two belt conveyors over bunker, each 36 in. wide, capacity each per 

hr., tons 

Feed pumps, 3 Jeansville, centrifugal, turbine-driven, all-bronze 

casings. 

Feed pump capacity per sq. ft. heating surface, gal. per min 

Make-up water evaporator system capacity, lb. per hr 

Present make-up water evaporator capacity, per cent of hot-well 

supply (based on 60,000 kw. turbine capacity) 

Main open feed-water heaters: Cochran horizontal cylindrical; ca- 
pacity each, boiler horsepower 

Heater capacity per sq. ft. boiler-heating surface, boiler horsepower 

(34.5 lb. water per hr.) 

Heater capacity per rated horsepower capacity of boilers, boiler 

horsepower 

Heater capacity per lb. main-unit steam consumption (95,000 kw. @ 

10.25 lb. per kw-hr.), boiler horsepower 

Chimneys: Two steel-lined. 

Contractors, Lackawanna Steel Co. Builders, Merchants Iron 
Works, Chicago. 

Height above lower grate, ft. 192; height above upper grate, ft. . . 

Diameter at flue entrance, ft 

Diameter at top, ft 

Boilers per chimney 

Coal burned per sq. ft. chimney cross-sectional area at approx. 400 

per cent rating, lb. per hr 

Forced-draft fans. Green, radial flow; number of 

Capacity of each at 6 in. static pressure, 550 r.p.m., cu. ft. per min. 

(hp. 308) 

Capacity of each at 4^ in. static pressure, 430 r.p.m., cu. ft. per min. 

(hp. 153) 

Capacity of each at 3 in. static pressure, 336 r.p.m., cu. ft. per min. 

(hp. 67) 

Induced-draft fans: Buffalo Forge Co., number of 

Induced-draft fan capacity, each, with gas at 496 deg. fahr., 482 

r.p.m. cu. ft. per min. (hp. 130) 

Forced-draft fan capacity per sq. ft. grate, cu. ft. per min 

Induced-draft fan capacity per sq. ft. grate, cu. ft. per min 

Bunker coal storage over present 8 boilers, maximum tons 

Yard coal storage at plant, tons 



0.053 
30,000 



10,000 
0.53 



5.3 



0.0308 



185 

19 

19 

4 



185. 



210,000 

150,000 

100.000 
6 



120,500 

251 

288 

3,000 

50,000 



926 



STEAM POWER PLANT ENGINEERING 



TABLE 1Q4:. — Continued. 
Turbines. 

Three 20,000-kw. at 90 per cent power factor, installed; one 35,000-kw. 

on order. 

General Electric Co., single-cylinder, horizontal, speed, r.p.m 1,500 

Operating pressure, lb. abs 290 

Operating superheat, deg. fahr 275 

Performance guarantees: Operating conditions — 265 lb. abs. 250 

deg. fahr. superheat. 1-in. absolute pressure, 30-in. barometer, in condenser: 



Net Kw. Load of 
Generator. 


Lb. of Steam per 
Kw-hr. 


7,500 
10,000 
15,000 
20,000 


11.85 
11.05 
10.25 
10.60 



Note. For higher pressures and temperatures the following factors are used: 
1 per cent for each 15 lb. pressure for range of 25 lb. above or below normal; 1 per 
cent for each 11 deg. fahr. superheat for range of 25 deg. fahr. above or below normal. 

Blading material: First 2 and last 3 rows, nickel steel; interme- 
diate rows, monel metal and nickel bronze. 

Peripheral speed last rows of low-pressure blading', ft. per sec 717 

Total weight each 20,000-kw. machine, lb 540,000 

Weight of turbine per rated kw. capacity, lb 27 

Heaviest piece to be lifted by crane, tons 70 

Floor space occupied by each turbine, outside measurements, sq. ft. 830 

Turbine rated capacity per sq. ft. floor covered by turbine, kw 24 

Steam consumption of auxiliaries: 

At most economical load, lb. per hr., 6100; with 70-deg. fahr. 

circ. water, lb. per hr 10,500 

At fall load, lb. per hr., 9100; with 70 deg. fahr. circ. water, lb. 

per hr 13,500 

Exciters: Three, capacity each, kw 300 

Type Combination turbine and induction motor drive 

Builders Terry Turbine Co. 

Main unit capacity per kw. capacity of exciters, installed, kw 105.5 

Condensers. 

Builder Westinghouse Electric and Mfg. Co. 

Total tube surface per condenser, sq. ft 33,000 

Tube area per kw. turbine capacity served by condenser, sq. ft 1 . 65 

Tubes, 1 in. O. D., composition, Muntz metal (60 per cent copper, 40 per cent zinc) 
Chief guarantee: With 70 deg. circulating water, pressure in con- 
denser, lb. abs 1 . 33 

Circulating pumps, capacity each, gal. per min 25,000 

Builder, Westinghouse Electric and Mfg. Co., type, double suc- 
tion, centrifugal. 

Diameter discharge pipe, in 42 

Circulating-water pumping capacity per lb. steam condensed at 

consumption of 10.6 lb. per kw-hr., lb 118 

Circulating pumps (two per condenser) , capacity each, gal. per min . 25,000 
Cross-sectional area each intake and each discharge tunnel for each 

unit, sq. ft 30 

Intake, tunnel area per 1000 gal. per min. circulating-water pump- 
ing capacity, sq. ft 0.6 

Screens at circulating water intake Wire mesh, stationary 

Dry-vacuum pump, type Le Blanc 

Hot-well pump: Centrifugal; builder, Worthington; size in 4 



TYPICAL CENTRAL STATIONS 927 
TABLE 165. 

COMMONWEALTH EDISON CO., NORTHWEST, UNIT No. 3 — GENERAL DATA. 

Turbine. 

Maker General Electric Co. 

Type Horizontal compound 

Capacity, hp 45,000 

Number single stages, h-p. element 10 

Number of double stages, 1-p. element 2 

Speed, r.p.m 1,500 

Condenser. 

Maker Wheeler Condenser and Engineering Co. 

Number of tubes 1 1,000 

Size of tubes, in 1 

Surface in condenser, sq. f t 50,000 

Surface per kilowatt of generator rating, .sq. ft 1 . 67 

Capacity, lb. of steam per lir 360,000 

Steam condensed per square foot of surface, lb 7.2 

Weight of condenser, empty, tons 176 

Weight of cooling water in condenser, tons 66 . 5 

Circulating-pump capacity, gal. per min 52,000 

Circulating-pump capacity, lb. per hr 26,000,000 

Cooling water per pound of steam, lb 72 

Condensate pump, gal. per min 1,200 

Generator. 

Maker General Electric Co. 

Capacity, rated 30,000 

Voltage 9,000 

Frequency, cycles 25 

Speed, r.p.m 1,500 

Number field poles 2 

Length complete unit, overall ft 59 . 5 

Width, ft 18.33 

Floor area cover, sq. ft 1,091 

Area per kilowatt of generator rating, sq. ft 0.036 

Exciter voltage 220 

Boilers. 

Maker Babcock & Wilcox Co . 

Type Cross-drum, water-tube 

Pressure, lb. per sq. in. gauge 230 

Superheat, deg. fahr 200 

Temperature of steam, deg. fahr 600 

Number of boilers in unit 5 

Number of tubes per boiler - 588 

Diameter of tubes, in 4 

Length of tubes, ft 18 

Steam-making surface in boiler, sq. f t 12,200 

Stokers per boiler 2 

Type of stoker B. & W. chain-grate 

Active area of two stokers, sq. ft 273 

Ratio grate area to boiler-heating surface 1 to 45 

Per 1000 sq. ft. of boiler-heating surface: 

Connected grate area, sq. ft 22 . 3 

Stack area, sq. ft 4.17 

Economizer surface, sq, ft 538. 2 

Capacity of each boiler, lb. steam per hr 85,000 

Evaporation per sn, ft. of heating surface, lb , , 7 



928 STEAM POWER PLANT ENGINEERING 



TABLE 165. — Continued. 

Coal capacity of each boiler, lb. per hr 12,600 

Coal per square foot of grate, lb 46 

Size of steam main to turbine, in 20 

Boiler-feed pumps : 

Maker Henry R. Worthington 

Type Turbine-driven, three-stage, double-suction impeller 

Capacity, lb. per hr 450,000 

Speed, r.p.m 2,500 

Water temperatures: 

In feed-water heater, deg. fahr. 100-120 

Leaving economizer and entering boiler, deg. fahr 270 

Economizers. 

Maker B. F. Sturtevant Co. 

Type High-pressure 

Number of economizers 5 

Number of tubes in each 456 

Length of tubes, ft 12 

Heating surface in tubes 6,566 

Maker B. F. Sturtevant Co. 

Type Multivane 

Capacity, cubic feet hot gases per min 90,000 

Horsepower of motor . 100 

Draft in boiler uptake, in. of water 2.4 

Height of stack above boiler-room floor, ft 250 

Diameter of stack, inside, ft 18 



ii 



CHAPTER XXI 

A TYPICAL MODERN ISOLATED STATION* 
Bleeder Turbines and Condenser System 

439. The new power plant of the W. H. McElwain Company at 
Manchester, N. H., is an excellent example of current practice in 
generation of power by steam for industrial purposes. 

General Arrangement. — General arrangement of the boiler and 
engine rooms is shown in plan in Fig. 613. At the present time there 
have been installed three 300-horsepower water-tube boilers and one 
1000-kilowatt turbo-generator outfit. The boiler room contains suffi- 
cient space for a fourth 300-horsepower unit, as indicated by dotted 
lines. The completed plant will include duplicates of the two batteries 
shown, making a total of 2400 horsepower. The future boilers will 
face those already installed, the building being extended for this pur- 
pose, and the firing space shown will be common to both sections. 

The chimney, which is 176 feet in height, with a flue 9 feet in diam- 
eter, is designed with reference to the final capacity of the plant. In 
the engine room, at the right, is shown space for two additional gener- 
ating units, which provide for an ultimate capacity of 3000 kilowatts. 
Sectional elevations, showing the boilers, turbines, and the various 
auxihary equipment and their connections, are illustrated in Figs. 612, 
613, and 614. 

Boilers. — Present equipment consists of three Babcock and Wilcox 
water-tube boilers, each containing 2972 square feet of heating surface 
and about 50 square feet of grate surface. The heating surface is made 
up of two steam drums, tubes, and mud drum, and a superheater of 
the form shown in Fig. 614. 

Each boiler contains 144 4-inch tubes, 18 feet in length, made up in 
12 sections of 12 tubes each, and 2 steam drums, 3 feet in diameter by 
20 feet 4 inches in length. The superheaters each contain approxi- 
mately 372 square feet of surface, which is 12^ per cent of the heating 
surface of the boiler, and are designed to give 100 degrees superheat 
when the boilers are operated at their normal rating of 300 horsepower. 
The proportions of all parts are designed for a working pressure of 160 
pounds per square inch and the safety valves are set at that point. 

* From the Practical Engineer, Chicago, July 1, 1912. 
929 



930 



STEAM POWER PLANT ENGINEERING 





r^ 



@= 



:zfev 



to 3 

- 7> 



A TYPICAL MODERN ISOLATED STATION 



931 




932 



STEAM POWER PLANT ENGINEERING 




A TYPICAL MODERN ISOLATED STATION 933 

Each boiler is provided with a water column fitted with high and low 
water alarm, try cocks, and gauge glass with special device for shutting 
off in case of breakage. Also 3|-inch lock pop-safety valve, and 12- 
inch steam gauge reading to 300 pounds pressure. The feed pipes are 2 
inches in diameter, provided with both check and gate valves, the latter 
having special extension handles. The blow-off connections are of 
2T}-inch extra heavy pipe, and are each provided with two blow-off 
valves of special design. 

Boiler settings are of hard-burned brick, laid in cement mortar, 
consisting of 1 part cement to 3 of sand, up to the level of the grates, 
and in lime mortar above that point. All parts of the furnaces and 
setting exposed to the fire are lined with firebrick laid in fire clay. The 
furnaces are of the '^ Dutch oven" type as shown in Fig. 614. 

Smoke Connections. — Location of the main smoke flue is best 
shown in Fig. 613. It is 4 feet 9 inches by 7 feet 6 inches in size and 
constructed of No. 10 black iron. It is stiffened with angle-iron 
braces and supported from the roof. The uptake from each boiler 
is provided with an adjusting damper for hand manipulation from 
the floor level. 

A balanced damper is located in the main flue at the point indicated, 
and operated by an automatic regulator of the hydraulic type. An 
interesting detail in connection with this work is the method of attach- 
ing the covering to the lower side of the flue so that it will not sag or 
peel off. This consists of cross pieces of 1-inch tee-bars placed 24 inches 
apart and riveted to the flue. The projecting flanges of these bars are 
drilled at frequent intervals and wires strung through, to which the 
covering is attached. 

Handling of Fuel and Ash. — Coal is brought to the fire room by 
cars running on a special track, as shown in Fig. 613. This track passes 
over platform scales just inside the building, where each load may be 
weighed as it is brought in. The track is double within the fire room 
so that the cars may pass, and also to furnish storage space for both 
coal and ash cars when so desired. 

The arrangement for the removal of ash is best shown in Figs. 613 
and 614. A dumping chute is provided in the bottom of each ashpit 
and at such an elevation that a car may be run underneath it as indi- 
cated. When filled, they are pushed to the ash lift (see Fig. 613) where 
they are raised to the boiler-room level and run out on the coal track 
for disposal. Combustible waste from the factory is })rought through 
a 36-inch pipe to a collector placed in the upper part of the boiler 
room, as shown in Fig. 614, and fed into the furnaces as there 
indicated. 



934 



STEAM POWER PLANT ENGINEERING 



Tvrbine and Generator. — The turbo-generator unit is one of the 
Westinghouse make, of 1000-kilowatt capacity, and equipped with an 
automatic bleeder connection and constant-pressure valve. It is 6 feet 
6 inches in width by 24 feet 8 inches in length and weighs approxi- 
mately 79,000 pounds. It is of the regular Westinghouse-Parsons 
type, the most interesting feature being the bleeder attachment which 
adapts it for use in combined power and heating plants. An impor- 
tant requirement for the economical operation of the ordinary steam 
turbine is the maintenance of a high vacuum at the exhaust end, 
which, of course, prevents the utilization of exhaust steam for heating 
purposes. 

The capacity of the turbine under different conditions is as follows: 
With a throttle pressure of 150 pounds per square inch (gauge), a 
vacuum of 28 inches, 100 degrees superheat, and a speed of 3600 
r.p.m., the normal capacity when condensing is 1500 b.hp. and the 
maximum 2250 b.hp. When running non-condensing with a back pres- 
sure not exceeding that of the atmosphere, the maximum capacity is 
1500 b.hp. 

It is interesting to note the probable steam economy of a turbine 
of this type when operating under varying loads, as expressed in 
the guarantee placed upon this machine, which is as follows: When 
operating under the above conditions, in connection with the gen- 
erator attached, the steam consumption per hour, including all leak- 
age and loss with the turbine, shall not exceed the quantities given 
below : 



Load, 


Power Factor, 


Kilowatts. 


Pounds Steam per 


Per Cent. 


Per Cent. 


Kilowatt-Hour. 


150 


80 


1500 


18.8 


125 


80 


1250 


18.3 


100 


80 


1000 


17.9 


75 


80 


750 


18.8 


50 


80 


500 


20.7 



When operating under the same general conditions, with 3 pounds 
gauge pressure at the bleeder connection, the steam consumption per 
hour shall not exceed the following at the loads indicated, when with- 
drawing the following amounts of steam through the bleeder connection: 



A TYPICAL MODERN ISOLATED STATION 



935 







Pounda of Steam 


Steam to Condenser. 


Load. 


Kilonattg. 






Per Cent. 














To Throttle. 


To Bleeder. 


Total. 


Kilowatts. 


150 


1500 


38,000 


18,600 


19,400 


12.9 






31,000 


10,000 


21,000 


14.0 


125 


1265 


38,000 


22,000 


16,000 


12.7 






36,300 


20,000 


16,300 


12.9 






29,200 


10,000 


19,200 


15.2 


100 


1000 


37,200 


30,000 


7,200 


7.2 






30,000 


20,000 


10,000 


10.0 






24,400 


10,000 


14,400 


14.4 


75 


716 


30,500 


30,000 


500 


0.7 






25,600 


20,000 


5,600 


7.8 






20,200 


10,000 


10,200 


14.2 


50 


469 


21,700 


21,700 





0.0 






20,600 


20,000 


600 


1.3 






16,000 


10,000 


6,000 


12.8 



Generator. — The generator is of the revolving-field type with inclosed 
frame, generating a 3-phase, 60-cycle, alternating current of 600 volts. 
The efficiency rating, with a power factor of 100 per cent, is as follows : 



Load, 
Per Cent. 


Efficiency, 
Per Cent. 


Load, 
Per Cent- 


Efficiency, 
Per Cent. 


50 

75 

100 


90.10 
93.00 
94.50 


125 
150 


95.50 
95.75 



Temperature rise based on its normal rating and a power factor of 
80 per cent, for periods of different length and for various loads, is 
given below : 



Load, 
Per Cent. 


Length of Run, 
Hours. 


Temperature Rise, 
Armature. 


Degs. F.. Field. 


100 

125 
150 


24 

24 
1 


72 

90 

108 


72 

90 

108 



The maximum conditions of continuous operation with a power factor 
of 80 per cent and for a room temperature of 77 deg. fahr. are as follows: 
Output, 1250 kilowatts (25 per cent overload). Rise in temperature: 

Armature, 90 deg. fahr. 
Field, 90 deg. fahr. 

Maximum temperature to which insulation can be subjected without 

^^j^^- Armature, 194 deg. fahr. 

Field, 302 deg. fahr. 



936 



STEAM POWER PLANT ENGINEERING 



There are two exciters provided, one being turbine driven and having 
a normal capacity of 25 kilowatts; the other motor driven, with a 
capacity of 40 kilowatts. The turbine is of the Westinghouse make, 
horizontal type, with a normal capacity of 38 b.hp. at a speed of 3500 
r.p.m. when running non-condensing, and a continuous overload capac- 
ity of 25 per cent. The steam requirements for this machine as re- 
gards temperature and pressure are the same as for the main turbine. 

The exciter is a direct-current machine with shunt winding, gener- 
ating a current of 125 volts at full load. 

Condensing Apparatus. — In connection with the main turbine a 
Westinghouse-Le Blanc jet condenser is used, and is shown in elevation 
in Figs. 614 and 615. This is designed to operate under a normal lift 
of 18 feet and takes its water supply from the intake tunnel as shown. 
When using injection water at a temperature of 70 deg. fahr. the fol- 
lowing results are guaranteed, with a water consumption not exceed- 
ing 724,000 pounds per hour: 



steam Condensed 

per Hour, 

Pounds. 


Vacuum Main- 
tained, Inches 
(Barometer, 30 Ins.) 


Steam Condensed 
per Hour, 
Pounds. 


Vacuum Main- 
tained, Inches 
(Barometer, 30 Ins.) 


10,350 
14,100 
18,000 


28.65 

28.44 
28.17 


19,950 
22,900 
30,000 


28.00 
27.80 
27.11 



The vacuum air pump is of the turbine type and is mounted upon 
the same shaft with the centrifugal ejector pump, both being driven by 
a steam turbine of 41 b.hp. running at 1500 r.p.m. under an atmos- 
pheric exhaust pressure. This piece of apparatus is shown at the base 
of the condenser in Figs. 614 and 615. 

High-pressure Piping System. — This includes all high-pressure piping 
in the boiler and engine rooms for the supply of turbines, pumps, etc., 
and for the supplementary supply to the heating system as may be 
needed. Pipe used for this purpose is full weight, wrought iron being 
used for sizes below 6 inches and open-hearth steel for larger sizes. 
The main drum at the rear of the boilers is of gun metal with nozzles 
cast in place. Expansion is provided for, so far as possible, by the use 
of sweep pipe bends and fittings of the long-turn pattern, all 2J-inch 
and larger fittings being of this design with flange joints. The high- 
pressure connections are shown in Figs. 613, 614, and 616. Starting 
at the boilers (Fig. 614), 6-inch leads are carried to a 12-inch drum 
supported on lower piers and rolls at the rear of the boilers. From here 
a 6-inch branch leads to the main turbine, and two branches of the 
same size to a 6-inch auxiliary main, running beneath the engine-room 



A TYPICAL MODERN ISOLATED STATION 



937 




938 STEAM POWER PLANT ENGINEERING 

floor, near to, and parallel with, the boiler-room wall. From this 
auxiliary main are taken the supplies to the various minor turbines and 
pumps, and also the branches leading to the low and intermediate- 
pressure system through reducing valves. The main drum is divided 
into two sections by means of a valve at the center, and each of these 
sections is connected with the auxihary drum as shown in Figs. 612 and 
616. The supplies to the various pumps are easily traced from Fig. 
616, also the connections with the 18-inch heating main and the inter- 
mediate-pressure line, leading to the factory through the tunnel leaving 
the building as indicated in the upper right-hand corner of the drawing. 

Exhaust System. — All low and intermediate pressure piping is full 
weight, sizes up to, and including, 12-inch being of wrought iron, while 
open-hearth steel is employed for the larger sizes. Standard-weight fit- 
tings are used for this work, those 6 inches and over being of the long- 
turn pattern. Flange joints are provided on all piping 2J inches and 
larger in diameter, the same as for high-pressure work. The exhaust 
piping is most clearly shown in Figs. 614, 615, and 616. Referring to 
Fig. 616 the 18-inch exhaust from the main turbine is shown as leading 
through a back-pressure valve into a 30-inch outboard line designed for 
the completed plant. This is clearly shown in elevation in Fig. 165. 
An 8-inch auxiliary exhaust connecting with the various pumps is 
shown in Fig. 615, parallel with, and below, the auxiliary high-pressure 
main already described. Steam from this enters the heating system 
through an oil separator. The 12-inch bleeder connection from the 
turbine leads to the 18-inch heating main and is shown in the same 
drawing, although more clearly in Fig. 616. 

Drainage. — The blow-off main from the boilers is carried directly to 
the river through a 4-inch cast-iron pipe. Drips from high-pressure 
piping are trapped to the main receiving tank and pumped back to the 
boilers. Exhaust drips, and all condensation containing oil, are trapped 
to a cast-iron sump tank located in the condenser pit, and, together 
with other drainage, are discharged by means of a water ejector. 

Water Supply, Feed Piping, etc. — Water for condensing and fire 
purposes is brought from the river through a cement conduit, a section 
of this, together with the 15-inch suction to the condenser, being shown 
in Figs. 614 and 615. The discharge from the condenser pump is into 
an 18-inch pipe leading to the river and shown in section in Fig. 614. 
Water pressure for fire protection is furnished by an 18 by 10 by 12-inch 
Underwriters' fire pump of 1000 gallons capacity, placed in the con- 
denser pit; this is shown in elevation in Fig. 615 and in plan in Fig. 616 
and takes its supply from the intake tunnel as there shown. 

The house tank and boilers have two sources of supply, one directly 



I 



A TYPICAL MODERN ISOLATED STATION 



939 






/Ti l 




J»s*° 



SZZil 



HI 



[HU 



«s 



;-^ 



Ot=ti^ 



03^^ s 



B 



fi 



i;=a 



JTq. 




s ET 



J7^ 



^ 






940 



STEAM POWER PLANT ENGINEERING 



from the city mains and the other from the intake tunnel. There is 
also a tank arrangement whereby water may be drawn from the dis- 
charge pipe of the condenser pump. 

These various lines are shown in Fig. 617. A 6-inch connection from 
the city main enters as shown at the upper part of the drawing, toward 
the left, and, after passing through a meter, branches are carried to the 




Plan of Condenser Piping. 



house tank, the receiving tank, the boilers, and to the priming pipes of 
the condenser and vacuum pumps. 

The second source of supply, that from the intake tunnel, requires 
the use of two turbine-driven house pumps of the one-stage turbine 
type, located in the condenser pit as shown in Fig. 617. These pumps 
each have a capacity of 200 gallons per minute against a head of 150 
feet, and discharge into a line of piping having branches connecting 
with the house tank, receiving tank, and boiler-feed pipe. A filtering 



A TYPICAL MODERN ISOLATED STATION 941 

equipment is also provided, as shown in Fig. 617, and so connected 
that the water from this source may be purified if desired. 

Boiler Feed. — Feed fines connecting with the boilers are shown in 
Fig. 613. One of these supplies water either directly by city pressure 
or from the turbine house pumps. The other supply is from a pair 
of boiler-feed pumps connecting with a receiving tank located in the 
boiler room as shown. The feed pumps, two in number, are of the 
duplex, outside packed, pot-valve type, 8 by 5 by 10 inches in size. 
The tank is 4 feet in diameter by 6 feet in length, of f-inch iron plate, 
and is connected with both city pressure and the house pumps. Under 
ordinary working conditions the feed supply is first discharged into the 
tank and then pumped to the boilers through a heater of 1000-horse- 
power capacity located as shown in the drawing. 

Heating System. — Factory buildings are heated by direct radiation 
in the form of coils and cast-iron radiators as best suited to local con- 
ditions. The Webster system of circulation is employed, a pair of 
6 by 10 by 12-inch single-piston vacuum pumps being connected with 
the main return as shown in Fig. 617. These discharge into the re- 
ceiving tank in the boiler room, and the condensation is pumped back 
to the boilers with the fresh feed. 

Steam supply for the radiation has already been mentioned, coming 
principally through the bleeder connection from the main turbine, 
supplemented, when necessary, by live steam through a reducing valve. 

Insulation. — In general, tanks, heaters, etc., are covered with 85 
per cent magnesia blocks, finished with a plastic coat of the same 
material, the total thickness of the covering, when finished, being 2 
inches. In addition to this, tanks and heater are provided with a 
covering of 7-ounce canvas. The insulation on that portion of the 
smoke pipe which comes outside of the building is protected by a cov- 
ering of heavy sheet iron. Steam piping, both high and low pressure, 
is insulated with 85 per cent magnesia sectional covering. All cold- 
water piping, with the exception of the connections to the condenser, 
are covered with wool felt, having a lining of tarred paper. Pipe 
covering of all kinds is finished with a heavy canvas jacket and painted. 



CHAPTER XXII.— Supplementary 

PROPERTIES OF SATURATED AND SUPERHEATED STEAM 

440. General. — The thermal and physical properties of water vapor 
though based on experimental data permit of accurate mathematical 
formulation, but the equations involved are too complex and unwieldy 
for everyday use. Tables and graphical charts calculated and plotted 
from these laws offer a simple and accurate means of solving practically 
all steam] problems and recourse to thermodynamic analysis is seldom 
necessary. 

Several tables and graphical charts of the properties of saturated 
and superheated steam have been published and though the values 
given by the various authorities differ somewhat from each other the 
variation is neghgible for most engineering purposes. The recent 
tables of Peabody,* Marks and Davis, f and of GoodenoughJ embody 
the latest and most accurate researches and are most commonly used 
in engineering practice. These tables give the simultaneous physical 
and thermal properties of saturated and superheated steam for various 
pressures and temperatures. All three tables are practically identical 
in arrangement as far as saturated steam is concerned but differ some- 
what in the treatment of superheated steam. 

441. Notations. — It is to be regretted that there is no accepted 
standard set of symbols for designating the various properties of steam. 
The use of different notations for the same property as in the case with 
the tables under consideration leads to much confusion. In the fol- 
lowing discussion an attempt has been made to follow general practice 
rather than that of any particular author. 

442. Standard Units. — The mean B.t.u. or yJo of the heat required to 
raise one pound of water from 32 deg. to 212 deg. fahr. is the accepted 
standard heat unit in all recent works on thermodynamics. 

The mechanical equivalent of heat / may be taken for all engineering 
purposes as 

1 mean B.t.u. = 778 standard ft. lb. 

(Goodenough, J = 777M; Marks & Davis, / = 777.54.) 

The reciprocal of J or yf^ is generally designated by the letter A. 

* Steam and Entropy Tables, Peabody, John Wiley & Sons, 1909. 
t Steam Tables and Diagrams, Marks & Davis, Longmans Green & Co., 1912. 
I Properties of Steam and Ammonia, Goodenough, John Wiley & Sons, 1915. 

942 



Properties of saturated and superheated steam 943 

The value of the absolute zero has been variously given as ranging 
from 459.2 to 460.66 deg. fahr. below zero. The most generally- 
accepted value is 459.6. For all engineering purposes, the value 460 
degrees is sufficiently accurate. Temperatures referred to zero deg. 
fahr. are generally designated by t and absolute temperature by T. 

The normal pressure of the atmosphere or one standard atmosphere 
is taken as 29.921 inch of mercury at 32 deg. fahr., or 14.6963 pounds 
per square inch. For most purposes these values may be taken as 
30 inches of mercury at ordinary room temperature and 14.7 pounds 
per square inch, respectively. Steam pressure should always be stated 
in absolute terms and not ''gauge" since the atmospheric pressure 
varies within wide limits. Notations j) and P are commonly used to 
designate pressure but because of the various methods of measuring 
this property they should be qualified to this effect. In the following 
discussion p represents pounds per square inch absolute and P pounds 
per square foot absolute. 

443. Quality. — This term applies strictly to the per cent of vapor in 
a mixture of vapor and water or wet steam and is usually designated 
by X] thus a quahty of 95 signifies that 95 per cent of the total weight 
of the mixture is vapor. For saturated steam x = \. The quality of 
superheated steam is designated by the temperature of the vapor or 
the degree of superheat. The latter term refers to the difference 
between the actual temperature and that of saturated vapor of the 
same pressure. 

444. Temperature-Pressure Relation. Saturated Steam. — All prop- 
erties of saturated steam depend on temperature only. For any 
temperature there is a corresponding pressure, the relationship being 
determined from formulas based upon experimental data. A large 
number of formulas have been proposed to represent this relationship 
but the more exact equations are too cumbersome for everyday use. 
In Marks & Davis' steam tables the pressure-temperature relationship 
is based upon the following law: 

log p = 10.51535-4873.71 T-i -0.00405096 T+0.00000139296 T\ (306) 

Wet Steam. — The relation between pressure and temperature is the 
same for wet steam as for saturated since the quality does not affect 
the temperature. 

Superheated Steam. — The temperature of superheated steam is not 
dependent solely upon the pressure and some additional property is 
necessary to fix the relationship. 

445. Specific Volume. Saturated Steam. — The specific volume s of 
saturated steam or the number of cubic feet occupied by one pound, 



944 STEAM POWER PLANT ENGINEERING 

varies with the pressure and is equal to the sum of the original volume 
of one pound of water cr, and u the increase in volume during vapor- 
ization, thus: 

s = u + <j. (307) 

Goodenough's modification of Linde's equation is 

u = 0.59465 - - (l + 0.0513 f) ^, (308) 

log m = 10.825. 

Wet Steam. — The specific volume v of wet steam may be calculated 

as follows : 

V = xs + {1 - x) o- (309) 

= xu -\- (T. (310) 

s is given in all saturated steam tables, o- varies from 0.0161 cu. ft. 
per lb. at a pressure of 1 lb. per sq. in., absolute, to 0.02 cu. ft. at 300 lb. 
a is so small compared with s that it may be neglected for most pur- 
poses and the specific volume becomes v = xs. v may be taken directly 
from the volume-entropy chart. 

Superheated Steam. — The specific volume of superheated steam v^ is 
given in all superheated tables. The values in Goodenough's tables 
were calculated from equation (308) by substituting u = v' — a. 

Wm. J. Goudie (Engineering, July 1, 1901) gives the following simple 
rule for determining the specific volume which gives satisfactory re- 
sults for moderate degrees of superheat. 

v' = s(l + 0.0016 O, (311) 

in which 

s = specific volume of saturated steam, pound per cubic foot, 
f = degree of superheat. 

Tumlirz's formula is a simple and fairly accurate abridgment of 
equation (308) for moderate degrees of superheat but at higher temper- 
atures gives results too low. 

v' = 0.5962 ^ - 0.256. (312) 

P 

446. Heat of the Liquid. — The heat of the liquid q, B.t.u. per pound 
above 32 deg. fahr., is the amount added to water at 32 deg. fahr. in 
order to bring it to the temperature of vaporization, thus: 



»/49; 



cdT, (313) 



in which c = specific heat at constant pressure. 



I 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 945 

c varies with the cemperature, but the relationship does not permit 
of simple formulation. If Cm = mean specific heat for the temperature 

range, 

q = Cm{t- 32). (314) 

For many purposes it is sufficiently accurate to assume c,„ = 1, 
then q = t — ?t2. The relationship between t, c, and Cm is shown in 
Fig. 618 for a wide range in temperatures. 

The heat of the Hquid is manifestly constant for a given temperature 
whatever may be the condition of the steam. 



















































































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SPECIFIC HEATS 
OF WATER 




















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50 100 150 200 250 300 350 400 
Temperature Degrees Fahrenheit 

Fig. 618. Specific Heats of Water. 

447. Latent Heat of Vaporization. — The latent heat of vaporization r, 
B.t.u. per pound above 32 deg. fahr., is the amount of heat required 
to change the fluid from a hquid to vapor at the same temperature. 
The latent heat has been accurately determined by direct experiment 
from 32 degrees to 356 deg. fahr. and numerous formulas have been 
based upon the experiments for calculating this quantity. Good- 
enough's values are calculated from the Clapeyron relation: 



r = A{8 - u)T 



dp 
df' 



(315) 



A simple formula which gives accurate results from 32 degrees to 
400 deg. fahr. is 

r = 970.4 - 0.655 {t - 212) - 0.00045 {t - 212)^. (316) 

At higher temperatures Hennings' exponential formula as modified 
by Dr. Davis is perhaps more accurate than equation (316), 

r = 139 (689 - iY^'^K (317) 

The latent heat decreases with the increase in temperature until a 
temperature of approximately 706 deg. fahr. (corresponding pressure 



946 STEAM POWER PLANT ENGINEERING 

3200 lb. per sq. in.) is reached when its value becomes 0. This is 
called the critical temperature. 

Values of r are given in all saturated steam tables. 

Special interest attaches to the values of r at 212 deg. fahr. because 
of its common use in engine and boiler tests. The following values 
have been assigned to this quantity. 

Regnault 966.0 Marks and Davis 970.4 

Peabody 969.7 Smith 972.0 

Heck 971.2 Goodenough 971.6 

The correct value is probably quite close to 972.0. 
External Latent Heat. — During the heating of the liquid the change 
in volume is very small and may be neglected, hence the external work 
done is negligible and also practically all of the heat goes to increase 
the energy of the liquid. During vaporization, however, the volume 
changes from a to s. Since the pressure remains constant, the external 
work that must be done to pro"\dde for increase in volume is 

P {s - a) = Pu (318) 

and the corresponding heat or external latent heat is 

AP {s - a) = APu. (319) 

Internal Latent Heat. — The heat r added during vaporization is 

used in increasing the energy and is doing external work. Hence the 

differeiice, or internal latent heat p, B.t.u. per pound above 32 deg. fahr., 

p = r - APu, (320) 

is the heat required to do disgregation work. 

448. Total Heat or Heat Content. — The total heat of saturated steam 
X, B.t.u. per pound above 32 deg. fahr., is evidently the sum of the heat 
of the liquid and the heat of vaporization, or 

X = r + g (321) 

= p + APu + q. (322) 

The total heat of saturated steam may be calculated by means of 
the Davis formula : 

X = 1046.187 + 0.6077 t - 0.00055 t\ (323) 

The quantity {p -\- q) -j gives the increase in energy of the saturated 

vapor over that of the liquid at 32 deg. fahr. and is called the intrinsic 
energy. 

Wet Steam.. — If vaporization is not complete the heat content Hw 
B.t.u. per pound above 32 deg. fahr. may be expressed: 

Hy, = xr -\- q (324) 

= xp + AP xu -^ q, (325) 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 947 

Superheated Steam. — If heat is added at constant pressure after 
vaporization is completed, the vapor will be superheated, and the 
heat content Ha is 

H, = r + q-\-Cmt' (326) 

= X + CJ\ (327) 

in which 

Cm = mean specific heat of the superheated vapor at constant pres- 
sure, 
t^ = degree of superheat = ^aup. — 4at.. 

Goodenough gives the following formula for calculating the total 
heat of superheated steam, absolute temperature of the steam T^ deg. 
fahr. 

H, = 0.320 T, + 0.000063 T/ -'-^- ^aP (l +^0.0342 p^) 

+ 0.00333 p + 948.7, ' ' (328) 

log Cs = 10.791155. 

449. Specific Heat of Steam. Saturated Steam. — If the amount of heat 
required to raise the temperature of saturated steam one degree and 
still maintain a saturated condition is construed as the specific heat 
of saturated steam, then the quantity is negative, since heat must be 
abstracted to effect this result. 

Csat. = 0.35 - 0.000666 {t - 212) - j^- (329) 

Superheated Steam. — The true or instantaneous specific heat C of 
superheated steam at constant pressure is the amount required to 
increase the temperature of one pound one degree fahr. Goodenough's 
equation based on the experiment of Knoblauch and Jakob is 

C = 0.320 + 0.000126 T. + ^^ + ^^^ ^' \T^^ "^^ , (330) 
log C2 = 11.3936. 

The mean specific heat may be calculated from superheated steam 
tables as follows: 

C^= ^^"^^r^ - (331) 

The true and mean specific heat of superheated steam at constant 
pressure for a wide range in pressures and temperatures arc shown in 
Figs. 619 and 620. The curves are taken from Goodenough's ''Prin- 
ciples of Thermodynamics." 



948 



STEAM POWER PLANT ENGINEERING 



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Fig. 619. True Specific Heat of Superheated Steam. 



PJIOPERTIES OF SATURATED AND SUPERHEATED STEAM 949 



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Superheat, Dog. F. 
Fig. 620. Mean Specific Heat of Superheated Steam. 



950 



STEAM POWER PLANT ENGINEERING 



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O to O to OS 
OO OO cs cs o 
''^ ^i^ ^^ ^^ to 



o o o o o 

'^l^ t^ OS cs CO 
to »0 to CO CO 



o o o o o 

OS CO t^ O to 
CO b- b- 00 00 



O to O to o 
o cs to t^ o 

cs cs <M d CO 



'ipui aj'Bnbg jad spuno^j 'aanssaj^j a^npsqy 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 051 

450. Entropy. General. — No change in a system of bodies that takes 
place of itself can increase the available energy of the system. As a 
matter of fact the actual physical process is accompanied by frictional 
effects and the quantity of energy available for transformation into 
work is decreased. This decrease in available energy or increase in 
unavailable energy is given the name increase of entropy. Although 
the solution of all engineering problems involving thermodynamic 
changes can be obtained without employing entropy, still its use 
simplifies the calculation in much the same manner that logarithms 
facilitate complex numerical computations. Increase of entropy be- 
tween the absolute temperatures T2 and Ti may be expressed 
mathematically 

Increase of entropy = / -^ ? (332) 

in which dQ represents an infinitesimal amount of heat and T the 
absolute temperature at which it is added. 

Entropy of the Liquid. — The increase in entropy 6 due to heating 
one pound of liquid from 32 deg. fahr. to temperature T is 



r'^=f'''-f, (333) 



in which 



Ti = absolute temperature of the liquid = ^1 + 460, 
q = heat of the hquid above 32 deg. fahr., B.t.u. per pound, 
c = specific heat of water at temperature T. 

Since c varies with the temperature according to a rather complex 
law, the integration in equation (333) does not reduce to a simple 
form. For example, Goodenough's equation for the range 32 — 212 
deg. fahr. assumes the form 

d = 2.3023 log T + 0.0045775 log (^ + 4) - 0.00022609 T 

H- 0.00000012867 T^ - 6.28787. (334) 

If the value of the mean specific Cm is known for the given tem- 
perature range equation (333) reduces to the simple form 

e = Cm\0ge^- (335) 

Values of are found in all unabridged steam tables. 



952 STEAM POWER PLANT ENGINEERING 

Entropy of Vaporization. — Since the temperature at which vapor- 
ization takes place is constant the change of entropy experienced by 
the fluid during vaporization is 

n = f=^- (336) 

If vaporization is incomplete as in case of wet steam 

xr 
ny, = xn = jpf' (337) 

Entropy of Superheat. — The entropy change during superheating 
may be expressed 

ns= j -^p-^ (338) 

Tv = temperature of the vapor. 

If the value of the mean specific heat Cm for the temperature range 
Tv to Tg is known the integration of equation (338) reduces to the 
simple form 

n, = C^loge^- (339) 

Total Entropy of Saturated Steam. — The increase in entropy from 
liquid at 32 deg. fahr. to saturated vapor at temperature T is 



N = n-\-d = ~-\-d. (340) 



Total Entropy of Wet Steam. 



TV 

Ny, = xn^e = '^-\-e. (341) 

Total Entropy of Superheated Steam. 

iV = n + n3 + = ^ + C^loge ^ + (9. (342) 

Using Knoblauch and Jakob's values for the specific heat of superheated 
steam, Goodenough gives the following rule for calculating the total 
entropy of superheated steam 

Ns = 0.73683 log T, + 0.000126 T, - ^^^^ - 0.2535 log p 

_ C.pil+^0.0M2p) _^^^^^^ ' (3^3^ 

log C, = 10.69464. 

Tables 167 and 168 are abridged from Marks and Davis' "Steam 
Tables and Diagrams. " 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 953 



Density- 
Weight per 
Cubic Foot, 
Pounds. 


?~ 


000340 
000656 
000961 


001259 
001555 
001850 


002143 
002431 
002719 


00300 
00576 
00845 


01107 
01364 
01616 


01867 
02115 
02361 


02606 
02849 
03090 


o oo 


o o o 


oo o 


o oo 


O O o 


o o o 


O OO 


Specific 
Volume. 


-^ 


o o o 


o o o 


o o o> 


O »0 lO 


CO 05 
»0 CO 00 


CO t— CO 
lO CM CO 


OO O CO 
CO ^H CO 


2935 
1524 
1041 


t^ CO »o 


t^ CM t>» 

CO —1 CO 
Tj* Tj< oo 


CO CO oo 
CO t^ — 1 

CO ^ .-• 


O CO ^ 
Oi t^ t^ 


CO t^ c^ 

>0 Tt< ■<*< 


oo lO CM 
CO CO CO 


It 


+ 
1. 


1728 
1127 
0775 


CO CO 00 

m (n '^ 
o o o 


CO CM CO 
lO -<^ -"^ 


-*' o oo 
lO oo •«*< 

t^ — 1 00 
Oi 05 00 


-* CM lO 

^ CO oo 

CO -^ CM 

00 oo 00 


^ CO oo 

CO lO lO 

.-H O OS 

00 00 !>. 


■* r- t-- 

t^ O CM 
oo t^ t^ 
t^ t^ t^ 


(M (M (M 


CM CM CM 


■ 


. 


. 


. 














Entropy 
of the 
Vapor. 


j.1^. 


1666 
0704 
0135 


9730 
9413 
9155 


CO t^ oo 
CO ■* t^ 

05 t^ lO 

00 00 00 


l^ w O 
CM CO ><*< 
-<*<'* 00 
oo t^ CO 


CO -* ■<*< 

^ 00 ^ 
rt< O 00 
CO CO lO 


CM O CM 

OO OO O 
lO CO CM 
lO lO lO 


CM »0 O 
-<*< 05 CO 

o oo t^ 

lO -<*< ■<*< 


(N CM (N 



























>» _: 

O, O "O 

8 5 3 

c <- .5* 



CM CO o o CO c:s 

CO CM ■* O CM CM 

O -^ CO OU 05 O 

O O O O O 1-H 



t^ lO lO 


t^ Oi oo 


OO OO r-H 


CTi CO CO 


CM CM t^ 


T-H 0> CO 


CM -* O 


05 -<tl t^ 


t-- t^ lO 


CO O CO 


!-< r-^ CM 


CO t^ O 


^ CO lo 


lO CO t^ 


oo 0> Ci 




.-H — 1 CM 


CM CM CM 


CM CM CM 


CM CM CM 



OOO OOO OOO OOO OOO OOO ooo 



•^lO^ lOCOO "^O-^ t^COOO OOOO "^OCO 1— liOOO 



OJ P; o <U O 



COCMt^ Tt^COOS t^OOCM OSt^Tji C0-*0 -^CM"** 050000 

b^lOt^ Clt^CO C0t--»O CMCOCO OOCMt^ CMOO"^ Ot^-^ 

r-Hoos 050000 oor^t^ t^iO"^ cococm cm— <-^ <— lOO 

0005 050505 Oi Oi Oi 050505 Oi <^ Oi Oi (^ Oi 0> Oi 05 



t>» 00 "^ 


CO CO ^ 


CO CM Oi 


"^ O CO 


lO lO t^ 


lO o ^ 


T-l OS lO 


Tt< CM t^ 

t^ 00 OO 
OOO 


T-H T*< t^ 


OS 1—1 CM 

OS o o 

O 1-1 --• 


"* to 1-1 
O — < CM 


CO O CO 
CM CO CO 


CO OS 1-1 

CO CO '"ji 


CO -"f CO 

M^ -rjl Tji 




t^ CO CO 


CO O OO 


^ t^ CO 


CO o CO 


t^ CO oo 


00 CM O 


O CM CO 



^M 1— < lO <0 t^ CO 
I^ CO lO lO -<*< >* 
OOO OOO 



^ 00 CO 

-<JH CO CO 
OOO 



■* ^ CM 
CO CM 1-1 
OOO 



lOO^O 1— lOOiO CMOSCO 

ooos osoooo oor^t^ 
OOOs OsosOi Ososos 






< f^ o 



COiOOS 1— lOOCM OOCOCO COiOCM i— lOOCO lOCOr^ CMiOCO 
Oi— 11^ OSCOCO 1— (-^CO OOi— 1»0 OCMO OOOOCM CMl^OS 



lO CO "<*< 

CO lO CO 


CM OS lO 
t^ t^ OO 


o -^ oo 
OS OS OS 


i-( CO 1-1 
O CM "* 


CO CM O 

lO CO t^ 


CO CM 00 

t— oo OO 


CO r- ^ 

OS crs o 


-H CM CO 
OOO 


•^ »0 CO 

dod 


t- oo OS 
OOO 


^ CM CO 


"^ »0 CO 


t- oo OS 


0-HC« 



954 



STEAM POWER PLANT ENGINEERING 






CO coco 

CO ^ t^ 

CO CO CO 

ooo 



coo 
ooc* 

CCOiy-* 
CO -* CO 
OOO 



00 1-1 Cv3 
(N "^ to 



lO to lO 
COl>.00 
O'-iC^ 



'^COfM 

o o i— I 

CO to CO 



(M(MCO 



(N (M (M 



OOO ooo ooo ooo ooo ooo ooo 



si 



05 

otocq 

00 CO '^l 



CiO0l>. t^COCO lOtOlO Tj^ Tfl rJH 



+ 



^ '^ to 
CO O CO 
CO CO to 



O O CO 
'^^(MCO 
to CO 1-1 



i-HOOi-l 

OiCO CO 
CO CO CO 



lO i-H to 
CO OOO 
CO to to 
CO CO CO 



C<1 00 1^ 

CO coo 

■<*l CO CO 
<X) CO CO 



(MO-H 
to O to 
(M C^ i-H 
CO CO CO 



CO CO CO 






»-l&i 



C5C0 t^ 

coc^-* 

CO tOTti 



-* C3iCO 

TtHCOCO 



,-( CO'* 

CO ooo 

COCOIM 



'^OOCi 

TjH coo 

""*CO 



^i 



O'* CO 

CO CO a> 

i-lOOO 

cq (M—i 



00 lOi-H 
t^CO CO 

t^co to 



CO CO t^ 
■^ CO(N 



c§°^ 



to i-i 00 
(M 00 1-1 



00 
. . OOi-l 
OO^ 
CO CO CO 



CO tO(M 
CO to CO 
1—1 CO to 
coco CO 



ooo o 

00 COC^ 
CO 00 Oi 
CO CO CO 



1-1 CO o 

(M 1-1 as 

O 1-^ 1— I 



(M Tt^ ^ 
(MCO^ 



•^ to o 

ir^co 05 
T^ lO to 



TjH "Tf CO 

-^ O Tti 
CO COt^ 



ooo ooo ooo ooo ooo ooo ooo 



c — 



03 «.„ titi 

a> > o ® o 

w 



(M CO Oi 


05 CO CO 


(N05C0 


1-1 coos 


1^00 1-1 


toi>. o 


(N tOt^ 




F2^[2 


^^f^ 


^?2g2 


^^8 


§Soo 


00 00 00 



® > O <» O 

«2, «^ 



OCOCO 000000 O i-< 05 C0O(N CO '^ CO 



(M Oil>. CO toco 

o oi as oi 00 1^ 

ooo 00 00 0000 



0<N to 

coco to 

00 00 00 



OiO O to 1-1 t^ 
to Tfi 'Tt^ coco (M 
OOOO 00 OOOOOO 



C0O5CO co< 
OOOOOO 00( 



O'^Tt^ 


t^<M "^ 


O 00-^ 


COCO'* 


oiooo 


00 CO'* 


r^Tt^CO 


00 05 O 

^Tt* to 

1— 1 »— 1 1— 1 


OCOO 
to to CO 

T— ( 1— 1 r-H 


SS8 

1-H I— 1 1— 1 


1-1 CO to 
t^ t^ t^ 


t^OO 05 


00 OOOO 


00^00 



w> 



(NO-* t>00 1-I05CO (NtOO 



^ i-iO 

l^ t^ t^ 

05 



00(M 
CO CO to 
05 



OOCO 05 

C^ CQ 1-1 

OOO 



oocq 

'*' 1-Ht^ 

ooo 



t> CO 1— ' ooo 



CO Ol>. 

ooo 

O OOO 



COOOO 
O O 00 
00 00 00 



Ot-htJI 

i-!coo6 



OOOi-i "<*tHCO i-^t^O "^OCO 



00 t^CO 
1-1 C^ CO 



t^C^ CO 

t^ 00 00 



to toco 
OTt^'od 

Oi 0> Oi 






iOO(M 



99'"! 

cooo o 

i-< C^ '* 



tooi-H t^oo cooco eoooo 



-* ^ t^ 

1^00 00 
(M (MCq 



<N 00(N 

ooo 

(M(M CO 



t^(McO 

CO CO CO 



O'^t^ 
CO CO CO 



-2 aJ G. © . 



1> 

CO "* '^ 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 055 



2365 
2472 
2577 


2683 
2791 
2897 


3002 
3107 
3213 


3320 
3425 
3529 


CO coco 


00 <M t>. 

rt< to«0 

O O 1— 1 
CO -^ '* 


4262 

437 

447 


t^QQOO 

toco t^ 

Tti ■* -^ 


OOO 
00 0(M 
■^ "^ to 


^ CO to 
to to CD 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


III 


coco (M 
(MOO to 
l:^lOrJ^ 


^05C^ 

coc^^ 


(M O-^ 

ogo8 


CO tOC^ 

to t^ o 

t-cDCD 


Sll 


cDOt^ 


t^OO ^ 

00 coo 


lii 


lis 


Ti<TtlCO 


CO COCO 


CO CO CO 


CO(M(M 


(M(M(M 


(M(M(M 


(M<N(M 


(M(M(M 


(M(Mt-( 


r-t^ ,-i 


.5980 
.6942 
.5907 


ill 


to to to 


.5692 
.5664 
.5639 


.5615 
.5590 
.5567 


§§| 


CD CO CD 

to to to 


CD 00 O 
1— 1 o t^ 

to to to 


CD^ O 
CO CO CO 
to to to 


.5276 
.5199 
.5129 




.1191 
.1108 
.1030 


.0954 

.0880 
.0809 


.0742 
.0675 
.0612 


.0550 
.0489 
.0431 


.0376 
.0321 

.0268 


.0215 
.0164 
.0114 


.0066 
.0019 
.9973 


ill 


§SJ2 


.9600 
.9419 
.9251 


1-1 rH 1-H 


1— 1 1— 1 T-l 


r— 1 T-H 1— 1 


i-H 1— 1 1— ( 


"" 


"" 


1-1 rH O 


Sd>d> 


odd 


OOO 


05Tt< !>. 

?2S§So 


Tt< "* "^ 


5035 
5072 
5107 


5142 
5175 
5208 


6625 
6925 
6825 


5328 
5356 

5384 


Ot^CO 


5488 
5513 
5538 


ill 


CDOOO 
t^OOl^ 

to to to 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


OOO 


O5i-<C0 


to CD 00 


05 O (M 


CO »0 CD 


l^ 00 o 


0^(M 


CO'-t^ to 


to CO t^ 


t^OO O 


T-H CO CO 


00 00 00 


(M(NCN 

00 00 00 


00 00 00 


S8S8S§ 


^s§s§ 


00 00^ 


00 00 00 


"^ -^ Ttl 

00 00 00 


oo^c» 


^^s 


C0t>.05 


(NCOO 


too I> 


-«*T-^00 


COiO-^ 


Tt^TjHTf< 


rt^iOCO 


OOOCO 


tooc* 


0(M OC 


iig 


to (NO 

05 05(0 


r^to(N 

00 00 00 


OOO to 


CO rHO 


§ii 


t^ t->- !>• 


t>- 1>- 1- 


!> !>• !>. 


t^ l:^ t^ 


(MOOO 


COCOO 


CD (MOO 


rt^OtO 


OTt<0 


Tt^ ooco 


1>^ to 


OOfMcD 


0(M O 


to 00 rH 


ggg 


OO c:5 05 


05 05 (35 


05 05 05 


gg§ 


§l§ 


§§§ 


§11 


§11 


§11 




(M lOOO 


(Mt^CO 


05CD -^ 


(M OOO 


00t^I> 


00 OOO 


0(M^ 


COO(M 


rt^oOtO 


CO CO CO 


tO(N05 

00 00 00 


t^rt<<M 
00 00 00 


ill 


SSI 


sis 


ill 


lOCO T-l 


i§i 


ill 


00 OC' 00 


OtOO 


CO to CO 


t^cO "* 


(M OCO 


(M l^(M 


cDOTt< 


t-O^ 


(M^'*< 


to »o-t< 


(M <M t^ 


ill 


<NiO00 

CO CO CO 


ii^ 


iWM 


CO CO CO 


CO CO CO 


(Mrt^ t^ 

»o to »o 

CO CO CO 


ill 


COCOt^ 
CO CO CO 


to -^ (M 
t^ 00 O 
CO CO CO 


TjHOOrH 


CO -^ ■* 


CO 1-H 00 


lOOcO 


O tooO 


^rt< CD 


OOOO 


OOO 


OOOrf 


,-HtO to 


iSi 


1— 1 -^ t-^ 


ill 


Scoco 
CO CO CO 


Hi 


CO CO CO 


%m 


ill 


Oi Oi C^ 

coco CO 




goto 


a^^ 


^^^ 


to toco 


§gK 


ooSo 


m 


228 

(M(M(M 


811 


OtOO 
to 1^ o 

(M(MCO 















956 



STEAM POWER PLANT ENGINEERING 



451. MoUier Diagram. — Steam tables are often accompanied by 
graphical charts that may be used to great advantage in the solution of 
thermodynamic problems. Fig. 621 gives a skeleton outline of the total 
heat-entropy diagram and Fig. 622 a reduced copy of the complete 
chart. The first conception of the heat-entropy chart is due to Dr. R. 
Mollier of Dresden, hence the name, Mollier Diagram. 

Referring to Fig. 621 abscissas represent total entropy and ordinates 
represent B.t.u. per pound. Vertical hues then indicate constant en- 
tropy and horizontal lines constant heat content. PiPi and P2P2 
represent lines of constant pressure and XiXi and X2X2 lines of con- 
stant quality. Evidently any point in the chart represents a fixed 




Total Entropy 

Fig. 621. Mollier Diagram — Skeleton Outline. 

condition of heat content, pressure, quality, and entropy as deter- 
mined by its location with respect to the different lines. Thus, point 
1 represents a pressure Pi as determined by the numerical value of 
line PiPi, quality Xi by its location on line XiXi, entropy A^i by its 
projection on the X axis and heat content Hi by its projection on the 
Y axis. 

In addition to the Mollier diagram the Marks and Davis tables 
include a total heat-pressure diagram which is of great assistance in 
the solution of problems involving ratios of expansion. 

The Ellenwood Charts (John Wiley & Sons, Pubhshers) have a 
much wider field of application than the diagrams mentioned above 
and afford a simple and accurate means of solving practically all ther- 
modynamic problems involving the use of the properties of steam. 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 957 



TABLE 168. 
PROPERTIES OF SUPERHEATED STEAM. 
Reproduced by Permission from Marks and Davis' " Steam Tables and Diagrams." 
(Copyright, 1909, by Longmans, Green & Co.) 



Pressure 




Satu- 


Degrees of 8uperhe.it. 




'ressure, 


PrMinrlt 




rated 
Steam. 






Po 11 n rl *j 


Absolute. 


50 


100 


150 


200 


250 


300 




Absolute. 




I 


162.3 


212.3 


262.3 


312.3 


362.3 


412.3 


462.3 


/ 




5 


V 


73.3 


79.7 


85.7 


91.8 


97.8 


103.8 


109.8 


V 


5 




h 


1130.5 


1153.5 


1176.4 


1199.5 


1222.5 


1245.6 


1268.7 


h 






t 


193.2 


243.2 


293.2 


343.2 


393.2 


443.2 


493.2 


i 




10 


V 


38.4 


41.5 


44.6 


47.7 


50.7 


53.7 


56.7 


V 


10 




h 


1143.1 


1166.3 


1189.5 


1212.7 


1236.0 


1259.3 


1282.5 


h 






t 


213.0 


263.0 


313.0 


363.0 


413.0 


463.0 


513.0 


t 




15 


V 


26.27 


28.40 


30.46 


32.50 


34.53 


36.56 


38.58 


V 


15 




h 


1150.7 


1174.2 


1197.6 


1221.0 


1244.4 


1267 . 7 


1291 . 1 


h 






t 


228.0 


278.0 


328.0 


378.0 


428.0 


478.0 


528.0 


t 




20 


V 


20.08 


21.69 


23.25 


24.80 


26.33 


27.85 


29.37 


V 


20 




h 


1156.2 


1179.9 


1203.5 


1227.1 


12.50.6 


1274.1 


1297.6 


h 






t 


240.1 


290.1 


340.1 


390.1 


440.1 


490.1 


540.1 


t 




25 


V 


16.30 


17.60 


18.86 


20.10 


21.32 


22.55 


23.77 


V 


25 




h 


1160.4 


1184.4 


1208.2 


1231.9 


1255.6 


1279.2 


1302.8 


h 






t 


250.4 


300.4 


350.4 


400.4 


450.4 


500.4 


550.4 


t 




30 


V 


13.74 


14.83 


15.89 


16.93 


17.97 


18.99 


20.00 


V 


30 




h 


1163.9 


1188.1 


1212.1 


1236.0 


1259.7 


1283.4 


1307 . 1 


h 






t 


259.3 


309.3 


359.3 


409.3 


459.3 


509.3 


559.3 


t 




35 


V 


11.89 


12.85 


13.75 


14.65 


15.54 


16.42 


17.30 


V 


35 




h 


1166.8 


1191.3 


1215.4 


1239.4 


1263.3 


1287 . 1 


1310.8 


h 






t 


267.3 


317.3 


367.2 


417.3 


467.3 


517.3 


567.3 


.t 




40 


V 


10.49 


11.33 


12.13 


12.93 


13.70 


14.48 


15.25 


V 


40 




h 


1169.4 


1194.0 


1218.4 


1242.4 


1266.4 


1290.3 


1314.1 


h 






t 


274.5 


324.5 


374.5 


424.5 


474.5 


524.5 


574.5 


t 




45 


V 


9.39 


10.14 


10.86 


11.57 


12.27 


12.96 


13.65 


V 


45 




h 


1171.6 


1196.6 


1221.0 


1245.2 


1269.3 


1293.2 


1317.0 


h 






t 


281.0 


331.0 


381.0 


431.0 


481.0 


531.0 


581.0 


t 




50 


V 


8.51 


9.19 


9.84 


10.48 


11.11 


11.74 


12.36 


V 


50 




h 


1173.6 


1198.8 


1223.4 


1247.7 


1271.8 


1295.8 


1319.7 


h 






t 


287.1 


337.1 


387.1 


437.1 


487.1 


537.1 


587.1 


t 




55 


V 


7.78 


8.40 


9.00 


9.59 


10.16 


10.73 


11.30 


V 


55 




h 


1175.4 


1200.8 


1225.6 


1250.0 


1274.2 


1298.1 


1322.0 


h 






t 


292.7 


342.7 


392.7 


442.7 


492.7 


542.7 


592.7 


t 




60 


V 


7.17 


7.75 


8.30 


8.84 


9.36 


9.89 


10.41 


V 


60 




h 


1177.0 


1202.6 


1227.6 


1252.1 


1276.4 


1300.4 


1324.3 


h 






t 


298.0 


348.0 


398.0 


448.0 


498.0 


548.0 


598.0 


t 




65 


V 


6.65 


7.20 


7.70 


8.20 


8.69 


9.17 


9.65 


V 


65 




h 


1178.5 


1204.4 


1229.5 


1254.0 


1278.4 


1302.4 


1326.4 


h 






t 


302.9 


352.9 


402.9 


452.9 


502.9 


552.9 


602.9 


t 




70 


V 


6.20 


6.71 


7.18 


7.65 


8.11 


8.56 


9.01 


V 


70 




h 


1179.8 


1205.9 


1231.2 


1255.8 


1280.2 


1304.3 


1328.3 


h 






I 


307.6 


357.6 


407.6 


457.6 


507.6 


557.6 


607.6 


i 




75 


V 


5.81 


6.28 


6.73 


7.17 


7.60 


8.02 


8.44 


V 


75 




h 


1181.1 


1207.5 


1232.8 


1257.5 


1282.0 


1306.1 


1330.1 


h 






I 


312.0 


362.0 


412.0 


462.0 


512.0 


562.0 


612,0 


t 




80 


V 


5.47 


5.92 


6.34 


6.75 


7.17 


7.56 


7.95 


V 


80 




h 


1182.3 


1208.8 


1234.3 


1259.0 


1283.6 


1307.8 


1331.9 


h 






t 


316.3 


366.3 


416.3 


466.3 


516.3 


566.3 


616.3 


I 




85 


V 


5.16 


5.59 


6.99 


6.38 


6.76 


7.14 


7.51 


V 


85 




h 


1183.4 


1210.2 


1235.8 


1260.6 


1285.2 


1309.4 


1333.5 


h 





t = Temperature, deg. fahr. 

V = Specific volume, in cubic feet, per pound. 

h = Total heat from water at .32 degrees, B.t.u. 



958 



STEAM POWER PLANT ENGINEERING 
TABLE 16S. — Continued. 



Pressure 




Satu- 


Degrees of Superheat. 


] 


Pressure, 


Pounds 
Absolute 




rated 
Steam. 






Pounds 
\.bsolute. 
















1 








50 


100 


150 


200 


250 


300 








t 


320.3 


370.3 


420.3 


470.3 


520.3 


570.3 


620.3 


t 




90 


V 


4.89 


5.29 


5.67 


6.04 


6.40 


6.76 


7.11 


V 


90 




h 


1184.4 


1211.4 


1237.2 


1262.0 


1286.6 


1310.8 


1334.9 


h 






t 


324.1 


374.1 


424.1 


474.1 


524.1 


574.1 


624.1 


t 




95 


V 


4.65 


5.03 


5.39 


5.74 


6.09 


6.43 


6.76 


V 


95 




h 


1185.4 


1212.6 


1238.4 


1263.4 


1288.1 


1312.3 


1336.4 


h 






t 


327.8 


377.8 


427.8 


477.8 


527.8 


577.8 


627.8 


t 




100 


V 


4.43 


4.79 


5.14 


5.47 


5.80 


6.12 


6.44 


V 


100 




h 


1186.3 


1213.8 


1239.7 


1264.7 


1289.4 


1313.6 


1337.8 


h 






t 


331.4 


381.4 


431.4 


481.4 


531.4 


581.4 


631.4 


t 




105 


V 


4.23 


4.58 


4.91 


5.23 


5.54 


5.85 


6.15 


V 


105 




h 


1187.2 


1214.9 


1240.8 


1265.9 


1290.6 


1314.9 


1339.1 


h 






t 


334.8 


384.8 


434.8 


484.8 


534.8 


584.8 


634.8 


t 




110 


V 


4.05 


4.38 


4.70 


5.01 


5.31 


5.61 


5.90 


V 


110 




h 


1188.0 


1215.9 


1242.0 


1267.1 


1291.9 


1316.2 


1340.4 


h 






t 


338.1 


388.1 


438.1 


488.1 


538.1 


588.1 


638.1 


t 




115 


V 


3.88 


4.20 


4.51 


4.81 


5.09 


5.38 


5.66 


V 


115 




h 


1188.8 


1216.9 


1243.1 


1268.2 


1293.0 


1317.3 


1341.5 


h 






t 


341.3 


391.3 


441.3 


491.3 


541.3 


591.3- 


641.3 


t 




120 


V 


3.73 


4.04 


4.33 


4.62 


4.89 


5.17 


5.44 


V 


120 




h 


1189.6 


1217.9 


1244.1 


1269.3 


1294.1 


1318.4 


1342.7 


h 






t 


344.4 


394.4 


444.4 


494.4 


544.4 


594.4 


644.4 


t 




125 


V 


3.58 


3.88 


4.17 


4.45 


4.71 


4.97 


5.23 


V 


125 




h 


1190.3 


1218.8 


1245.1 


1270.4 


1295.2 


1319.5 


1343.8 


h 






t 


347.4 


397.4 


447.4 


497.4 


547.4 


597.4 


647.4 


t 




130 


V 


3.45 


3.74 


4.02 


4.28 


4.54 


4.80 


5.05 


V 


130 




h 


1191.0 


1219.7 


1246.1 


1271.4 


1296.2 


1320.6 


1344.9 


h 






t 


350.3 


400.3 


450.3 


500.3 


550.3 


600.3 


650.3 


t 




135 


V 


3.33 


3.61 


3.88 


4.14 


4.38 


4.63 


4.87 


V 


135 




h 


1191.6 


1220.6 


1247.0 


1272.3 


1297.2 


1321.6 


1345.9 


h 






t 


353.1 


403.1 


453.1 


503.1 


553.1 


603.1 


653.1 


t 




140 


V 


3.22 


3.49 


3.75 


4.00 


4.24 


4.48 


4.71 


V 


140 




h 


1192.2 


1221.4 


1248.0 


1273.3 


1298.2 


1322.6 


1346.9 


h 






t 


355.8 


405.8 


455.8 


505.8 


555.8 


605.8 


655.8 


t 




145 


V 


3.12 


. 3.38 


3.63 


3.87 


4.10 


4.33 


4.56 


V 


145 




h 


1192.8 


1222.2 


1248.8 


1274.2 


1299.1 


1323.6 


1347.9 


h 






t 


358.5 


408.5 


458.5 


508.5 


558.5 


608.5 


658.5 


t 




150 


V 


3.01 


3.27 


3.51 


3.75 


3.97 


4.19 


4.41 


V 


150 




h 


1193.4 


1223.0 


1249.6 


1275.1 


1300.0 


1324.5 


1348.8 


h 






t 


361.0 


411.0 


461.0 


511.0 


561.0 


611.0 


661.0 


t 




155 


V 


2.92 


3.17 


3.41 


3.63 


3.85 


4.06 


4.28 


V 


155 




h 


1194.0 


1223.6 


1250.5 


1276.0 


1300.8 


1325.3 


1349.7 


h 






t 


363.6 


413.6 


463.6 


513.6 


563.6 


613.6 


663.6 


t 




160 


V 


2.83 


3.07 


3.30 


3.53 


3.74 


3.95 


4.15 


V 


160 




h 


1194.5 


1224.5 


1251.3 


1276.8 


1301.7 


1326.2 


1350.6 


h 






t 


366.0 


416.0 


466.0 


516.0 


566.0 


616.0 


666.0 


t 




165 


V 


2.75 


2.99 


3.21 


3.43 


3.64 


3.84 


4.04 


V 


165 




h 


1195.0 


1225.2 


1252.0 


1277.6 


1302.5 


1327.1 


1351.5 


h 






t 


368.5 


418.5 


468.5 


518.5 


568.5 


618.5 


668.5 


t 




170 


V 


2.68 


2.91 


3.12 


3.34 


3.54 


3.73 


3.92 


V 


170 




h 


1195.4 


1225.9 


1252.8 


1278.4 


1303.3 


1327.9 


1352.3 


h 





t = Temperature, deg. fahr. 

V = Specific volume, in cubic feet, per pound. 

h = Total heat from water at 32 degrees, B.t.u. 



PROPERTIES OF SATURATED AND SUPERHEATED STEAM 959 

TABLE 168. — Continued. 



Pre.ssure, 

Pound.s 

Absolute. 


' Degrees of Superheat. 
Satu- 1 


Pressure, 
Pounds 


rated 
Steam. 


50 


100 


150 


200 

570.8 


250 


300 


Ab.solute. 




t 


370.8 


420.8 


470.8 


520.8 


620.8 


670.8 


t 




175 


V 


2.60 


2.83 


3.04 


3.24 


3.44 


3.63 


3.82 


V 


175 




h 


1195.9 


1226.6 


1253.6 


1279.1 


1304.1 


1328.7 


1353.2 


h 






t 


373.1 


423.1 


473.1 


523.1 


573.1 


623.1 


673.1 


t 




180 


V 


2.53 


2.75 


2.96 


3.16 


3.35 


3.54 


3.72 


V 


180 




h 


1196.4 


1227.2 


1254.3 


1279.9 


1304.8 


1329.5 


1353.9 


h 






t 


375.4 


425.4 


475.4 


525.4 


575.4 


625.4 


675.4 


t 




185 


V 


2.47 


2.68 


2.89 


3.08 


3.27 


3.45 


3.63 


V 


185 




h 


1196.8 1227.9 | 


1255.0 


1280.6 


1305.6 


1330.2 


1354.7 


h 






t 


377.6 


427.6 


477.6 


527.6 


577.6 


627.6 


677.6 


t 




190 


V 


2.41 


2.62 


2.81 


3.00 


3.19 


3.37 


3.55 


V 


190 




h 


1197.3 


1228.6 


1255.7 


1281.3 


1306.3 


1330.9 


1355.5 


h 






t 


379.8 


429.8 


479.8 


529.8 


579.8 


629.8 


679.8 


t 




195 


V 


2.35 


2.55 


2.75 


2.93 


3.11 


3.29 


3.46 


V 


195 




h 


1197.7 


1229.2 


1256.4 


1282.0 


1307.0 


1331.6 


1356.2 


h 






t 


381.9 


431.9 


481.9 


531.9 


581.9 


631.9 


681.9 


t 




200 


V 


2.29 


2.49 


2.68 


2.86 


3.04 


3.21 


3.38 


V 


200 




h 


1198.1 


1229.8 


1257.1 


1282.6 


1307.7 


1332.4 


1357.0 


h 






t 


384.0 


434.0 


484.0 


534.0 


584.0 


634.0 


684.0 


t 




205 


V 


2.24 


2.44 


2.62 


2.80 


2.97 


3.14 


3.30 


V 


205 




h 


1198.5 


1230.4 


1257.7 


1283.3 


1308.3 


1333.0 


1357.7 


h 






t 


386.0 


436.0 


486.0 


536.0 


586.0 


636.0 


686.0 


t 




210 


V 


2.19 


2.38 


2.56 


2.74 


2.91 


3.07 


3.23 


V 


210 




h 


1198.8 


1231.0 


1258.4 


1284.0 


1309.0 


1333.7 


1358.4 


h 






t 


388.0 


438.0 


488.0 


538.0 


588.0 


638.0 


688.0 


t 




215 


V 


2.14 


2.33 


2.51 


2.68 


2.84 


3.00 


3.16 


V 


215 




h 


1199.2 


1231.6 


1259.0 


1284.6 


1309.7 


1334.4 


1359.1 


h 






t 


389.9 


439.9 


489.9 


539.9 


589.9 


639.9 


689.9 


t 




220 


V 


2.09 


2.28 


2.45 


2.62 


2.78 


2.94 


3.10 


V 


220 




h 


1199.6 


1232.2 


1259.6 


1285.2 


1310.3 


1335.1 


1359.8 


h 






t 


391.9 


441.9 


491.9 


541.9 


591.9 


641.9 


691.9 


I 




225 


V 


2.05 


2.23 


2.40 


2.57 


2.72 


2.88 


3.03 


V 


225 




h 


1199.9 


1232.7 


1260.2 


1285.9 


1310.9 


1335.7 


1360.3 


h 






t 


393.8 


443.8 


493.8 


543.8 


593.8 


643.8 


693.8 


t 




230 


V 


2.00 


2.18 


2.35 


2.51 


2.67 


2.82 


2.97 


V 


230 




h 


1200.2 


1233.2 


1260.7 


1286.5 


1311.6 


1336.3 


1361.0 


h 






i 


395.6 


445.6 


495.6 


545.6 


595.6 


645.6 


695.6 


t 




235 


V 


1.96 


2.14 


2.30 


2.46 


2.62 


2.77 


2.91 


V 


235 




h 


1200.6 


1233.8 


1261.4 


1287.1 


1312.2 


1337.0 


1361.7 


h 






t 


397.4 


447.4 


497.4 


547.4 


597.4 


647.4 


697.4 


t 




240 


V 


1.92 


2.09 


2.26 


2.42 


2.57 


2.71 


2.85 


V 


240 




h 


1200.9 


1234.3 


1261.9 


1287.6 


1312.8 


1337.6 


1362.3 


h 






t 


399.3 


449.3 


499.3 


549.3 


599.3 


649.3 


699.3 


I 




245 


V 


1.89 


2.05 


2.22 


2.37 


2.52 


2.66 


2.80 


V 


245 




h 


1201.2 


1234.8 


1262.5 


1288.2 


1313.3 


1338.2 


1362.9 


h 






t 


401.0 


451.0 


501.0 


551.0 


601.0 


651.0 


701.0 


I 




250 


V 


1.85 


2.02 


2.17 


2.33 


2.47 


2.61 


2.75 


V 


250 




h 


1201.5 


1235.4 


1263.0 


1288.8 


1313.9 


1338.8 


1363.5 


h 






t 


402.8 


452.8 


502.8 


552.8 


602.8 


652.8 


702.8 


i 




255 


V 


1.81 


1.98 


2.14 


2.28 


2.43 


2.56 


2.70 


V 


255 




h 1201.8 


1235.9 


1263.6 


1289.3 


1314.5 


1339.3 


1364.1 h 





t = Temperature, deg. fahr. 

V = Specific volume, in cubic feet, per pound. 

h = Total heat from water at 32 degrees, B.t.u. 



CHAPTER XXIII — Supplementary 

ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 

452. GeneraL — The laws governing the transformation of steam 
from one state to another form the basis of practically all thermodynamic 
analyses of the steam engine and turbine. The more common and 
important changes are 

(1) Isobaric or equal pressure. 

(2) Isovolumic or equal volume. 

(3) Isothermal or equal temperature. 

(4) Constant heat content. 

(5) Adiabatic or no external heat exchange. 

(6) Poly tropic. 

453. Isobaric or Equal Pressure Change. Saturated Vapor. — Since 
the temperature of wet or saturated steam is dependent on the pres- 
sures only, a constant pressure change of such material must also be a 
constant temperature one. Denoting the initial and final properties 
by subscripts 1 and 2 respectively: 

Initial volume Vi = XiSi + (1 — Xi) o-i = XiUi + ai. (344) 

Final volume V2 = X2S1 + (1 —X2) ai = X2U1 + o-i. (345) 

Change of volume z;2 — t^i = Ui {x2 — Xi). (346) 

External work W = Pi {v2 — Vi) = PiUi {x2 — Xi). (347) 

Change of energy = -j {x2 — Xi). (348) 

Heat absorbed = n {x2 — Xi). (349) 
Notations : 

.1 1, . , H = heat content above 32 dee, fahr., 

^ " 778' ^^^^' P^' ''^- '"'■ ^^^- B.t.u. per lb. 

P = lb. per sq. ft. abs. x = quality ^ = total heat of dry steam, B.t.u. per 

of wet steam. lb. 

s = specific volume of dry steam, ^ = latent heat of vaporization, B.t.u. 

lb. per cu. ft. . Per lb. 

V = specific volume of vapor, lb. p = internal latent heat, B.t.u. per 

per cu. ft. lb. 

a- = specific volume of water, lb. Q = heat of hquid, B.t.u. per lb. 

per cu. ft. ^ = entropy of the liquid. 

u = increase in volume during ^ = entropy of the vapor. 

evaporation, cu. ft. ^ = total entropy. 

t = deg. fahr. above zero. T = Prime marks indicate superheat. 

deg. fahr. abs. Subscripts 1, 2, w, s indicate, respec- 

Cm = mean specific heat of water. tively, initial condition, final condi- 

C = mean specific heat of super- tion, wet steam, and superheated 

heated steam. steam. 
960 



ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 961 

Example 77. At a pressure of 115 lb. per sq. in. absolute the vol- 
ume of one pound of vapor and liquid is increased 1 cu. ft. Required 
the change of quality, external work, increase of energy and heat ab- 
sorbed. 

From steam tables Si = 3.88; en = 0.0179; pi = 797.9; n = 879.8. 

Change of quality = x, - x, = '^^ = 3 gg \^^^^ = 0.259. 

External work = Pi fe - Vi) = 144 X 115 X 1 = 16,560 ft. lb. 

Change of energy = ^ (^2 - Xi) = 797.9 X 778 X 0.259 

= 160,778 ft. lb. 
Heat absorbed = n {x2 - Xi) = 879.8 X 0.259 = 227.79 B.t.u. 

Superheated Steam. — Let superheated steam change state at con- 
stant pressure pi from an initial temperature ti to a final temperature ^2. 

Change of volume = V2' — V\ . The values of v' corresponding to 
pressure pi and temperatures ^1 and t^ may be taken directly from 
steam tables or they may be calculated from equation (308). They 
may be approximated from equations (311) and (312). 

External work = Pi {v^! - Vi'). (350) 

Change of energy = (^^ - P.v^'^ - {^ - P,v,'^ • (351) 

= ^' ~ ^' - PM' - V,'). (352) 

Heat absorbed = H^' - H,', (353) 

Change of entropy = N2' - iV/. (354) 

Example 78. Using the data in the preceding example determine the 
various quantities, if the initial degree of superheat is 100 deg. fahr. 

From superheated steam tables for pi = 115 and ty = 438.1 (= 338.1 
+ 100) we find: v,' = 4.51; H,' = 1243.1; N,' = 1.6549. 

For P2 = Pi = 115 and V2 = (4.51 -h 1) = 5.51 we find by interpola- 
tion H2' = 1328.5; N2' = 1.7419; ^2' = 621.3. 

Increase of superheat = 12' — ti 

= 621.3 - 438.1 = 183.2 deg. fahr. 
External work = Pi (v2 — vi') 

= 144 X 115 X 1 = 16,560 ft. lb. 

Increase of energy = ^ ^ — Pi {v2 — Vi') 

= (1328.5 - 1243.1)778 - 16,560 
= 49,881 ft. lb. 
Increase of entropy = A/'/ — A^2' 

= 1.7419 - 1.6549 = 0.087. 
Heat absorbed = H2 — Hi' 

= 1328.5 ^ 1243.1 = 85.4 B.t.u. 



962 STEAM POWER PLANT ENGINEERING 

454. Isovolumic or Equal Volume Change. Saturated Steam. — Since 
the volumes Si and S2 are equal 

Si = S2 or XiUi + (Ti = X2li2 + 0-2. (355) 

External work = 0. (356) 

Heat absorbed = Xipi + gi — (X2P2 + ^2). (357) 

Example 79. A pound of mixture of vapor and liquid at 115 lb. per 
sq. in. absolute and quality 0.9 is cooled at constant volume to a pres- 
sure of 1 lb. per sq. in. absolute. Required the various properties 
at the final condition and the heat taken from the mixture. 
From steam tables: 

pi = 115, si = 3.88,(71 = 0.0179, 
Pi = 797.9, gi = 309, Ui = 1.103, 6, = 0.4877, 
P2 = 1, S2 = 333, 0-2 = 0.0161, p2 = 972.9, 
q2 = 69.8, 712 = 1.8427, 62 = 0.1327, 

-n,. 1 y. XiUi + 0-1 — 0-2 

Fmal quality X2 = 

^ 0.9(3.88 - 0.0179) + 0.0179 - 0.0161 

333 - 0.0161 
= 0.0105. 
Heat removed = Xipi -\- qi — {X2P2 + ^2) 

= 0.9 X 797.9 + 309 - (0.0105 X 972.9 + 69.8) 
= 947 B.t.u. 
Initial entropy Ni — XiUi + di 

= 0.9 X 1.103 + 0.4877 = 1.4804. 
Final entropy iV2 = 0:2^2 + 62 

= 0.0105 X 1.8427 + 0.1327 = 0.1520. 

Superheated Steam. — Since the final volume is equal to the initial , 
and both pressures and the initial temperature are known, the final 
temperature may be calculated from equation (308) or it may be taken 
directly or interpolated from the steam tables. 

Example 80. Using the data in the preceding problem determine the 
various factors if the initial degree of superheat is 100 deg. fahr. 

From steam tables for pi = 115 and ^1 = 338.1 + 100 = 438.1 we 
find: vi' = 4.51, i// = 1243.1, AT/ = 1.6549. 

s for 1 lb. per sq. in. absolute pressure = 333 cu. ft. but the given 
volume is 4.51 cu. ft. Therefore the steam is wet at the final condition. 

From steam tables for ^2 = 1 we find: 

P2 = 972.9, q2 = 69.8, n2 = 1.8427, $2 = 0.1327. 
Since the volumes are equal 

Vi = V2 

= X2U2 + a 2. 
Final quality X2 = — 

" 333 - 0.0161 " ^•^^^^' 



ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 963 

Heat removed = Hi — APiOi — (X2P2 + ^2) 

1 44. V 11^ 
= 1243.1 - jj^ X 4.51 - (0.0135 X 972.9 + 69.8) 

= 1065 B.t.u. 

Initial entropy (from steam tables) iV/ = 1.6549. 

Final entropy A^2 = ^2^2 + ^2 

= 0.0135 X 1.8427 + 0.1327 = 0.1575. 

455. Isothermal or Equal Temperature Change. Saturated Vapor. — 
Since the temperature of wet or saturated steam is dependent solely upon 
the pressure, an isothermal change is also isobaric, and the data in 
paragraph (458) is applicable to this change. 

Superheated Steam. — The properties at initial and final conditions 
may be calculated from equations of the properties of superheated 
steam or they may be taken directly from steam tables or charts. If 
wet or saturated steam expands isothermally into the superheated state 
the pressure must drop in order to maintain constant temperature. 
The relation between pressure, volume, and temperature for the super- 
heated state is given in equation (308). 

Example 81. One pound of steam at initial pressure 115 lb. per sq. 
in. absolute and superheat 100 deg. fahr. is expanded isothermally 
to a pressure of 1 lb. per sq. in. absolute. Required the various proper- 
ties at the final pressure, the heat absorbed during expansion and the 
external work done. 

From superheated steam tables for pi = 115 and ti = 338.1 + 100 
= 438.1 we find: Vi' = 4.51, Hi' = 1243.1, Ni' = 1.6549. 

For p2 = 1 and ^' = 438.1, V2' = 535, H2' = 1258.3, N2' = 2.1888. 

Final quality ^2' - 4 = 438.1 - 101.8 = 336.3 deg. superheat. 

Heat added during expansion = T2 {N2 — Ni'). 

= 898(2.1888- 1.6541) 
= 1378 B.t.u. 

(Note that the heat added is not equal to the difference in total 
heats since the isothermal is not a constant pressure line.) 



External work = 



Cpdv. (358) 



Since the temperature is constant dv may be obtained by differen- 
tiating equation (308). Substituting this value of dv in equation (358) 
and integrating we have, 

External work = 85.63 loge ^ ^- + 2.46 (pi^ - pA ~ (359) 

= 85.63 loge 898 ^ + 2.46 (ll5^ - 1^) ^^, 

= 368,000 ft. lb. (approx.) 
(log C = 10.8250.) 



964 STEAM POWER PLANT ENGINEERING 

456. Constant Heat Content. — Expansion from one pressure to a 
lower one with constant heat content is exemphfied in throttling or 
wire drawing. The energy utilized in imparting velocity to the fluid 
is all returned to the fluid at the lower pressure when the velocity is 
brought to zero and there are no radiation losses. 

For steam wet throughout expansion 

Xin -\- q\ = X2r2 -\- ^2. (360) 

For steam initially wet but dry at the lower pressure 

Xin -\- qi = X2. (361) 

For steam initially wet but superheated at the lower pressure 

xin + 51 = X2 + O2' = H2'. (362) 

For steam initially dry 

Xi = X2 + Cmk' = H2'. (363) 

For steam initially superheated 

Hi' = H2'. (364) 

Loss of available energy due to throttling or wire drawing 

Loss B.t.u. per lb. = T2 {N2 - Ni). (365) 

Example 82. One pound of steam at an initial pressure of 115 lb 
per sq. in. absolute is expanded through a throttling calorimeter to 
a pressure of 16 lb. per sq. in. absolute. If the temperature of the 
steam at the lower pressure is 256.3 deg. fahr. required the initial 
quality of the steam. 

From saturated steam tables: 

Pi = 115, n = 879.8, gi = 309, iVi = 1.5907. 
From superheated steam tables for p2 = 16 and ^2' = 256.3 we find: 

H2 = 1170.8, N2 = 1.7765, h (sat.) = 216.3, 
Xin -\- qi = H2, 
879.8 a^i + 309 = 1170.8, Xi = 0.98. 

Mollier diagram analysis. Fig. 622. From intersection of constant 
superheat line t2 = 40 ( = 256.3 — 216.3) and constant pressure hne 
P2 = 16 trace horizontally to constant pressure line pi = 115 and read 
from its intersection with the constant quality line, Xi = 0.98. 

Decrease of available energy = T2 {N2 — Ni) (366) 

= (216.3 + 460) (1.7765 - 1.5907) 
= 125.6 B.t.u. 

457. Adiabatic Change of State. — Since in an adiabatic change there 
is no heat added to or abstracted from the fluid the entropy remains 
constant. 



ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 965 

Steam wet throughout change of state 

iVi = A^2. (366a) 

XiUi + ^1 = X2n2 + 02. (367) 

^ + ^,=^ + ^, (368) 

For water only x = 0; for dry steam x = 1. 
Steam initially superheated but finally wet 

A^i' = N2. (369) 

Ni + Us = X2n2 + 62. (370) 

Steam superheated throughout change of state 

N,' = N2', . (371) 

N, + ns = N2 + n,(2), (372) 

^ + ^1 + C. log. ^ = ^^ + ^2 + [C^ log. ^] + 02. (373) 

Final Quality. Saturated Steam. — This quantity may be calculated 
directly from equations (366a) and (367). 

X, = -^^ (374) 

= (^ + «,-».)g- (375) 

If water only is present at the beginning of expansion substitute 
iVi = ^1 in equation (374). 

For initial qualities of X\ — 0.50 (approx.) or greater the final 
quahty X2 decreases as the expansion progresses, and for initial quali- 
ties of x^^ = 0.50 (approx.) or less the final quality increases. For 
initial quality Xi = 0.50 the final quality X2 remains practically 
constant. 

The final volume may be calculated as follows : 
Wet steam, V2 = X2U2 + (T2, (376) 

X2 as calculated from equations (367) and (370), 
Dry steam, V2 = S2. (377) 

Superheated Steam. — For superheat at the end of expansion the 
calculations involved in equation (373) are too cumbersome and un- 
wieldy and the Molher diagram may be used to advantage. 

Volume Change. — Superheated steam: the final volume V2 may be 
calculated from equation (3) by substituting for p the final pressure^ 
and for Ts the final temperature as calculated from equation (373). 
The final volume, however, may be taken directly from the pressure- 
entropy chart. 



966 STEAM POWER PLANT ENGINEERING 

External Work. — Since the heat added or subtracted is zero, the 
external work is equal to the change of intrinsic energy, or in general 

W = j [{H, - AP.vO - (H2 - AP2V2)]' (378) 

Steam initially wet 

W = 2 lixipi + qO - ix2P2 + 52)]. (379) 

Steam initially dry, substitute Xi = 1. 

Steam initially superheated but wet at end of expansion 

Tf = J [(^/ - AP,v,') - {X2P2 + 52)]. (380) 

Steam initially superheated but dry at end of expansion substitute 
X2 = 1. 

Steam superheated throughout expansion 

W = j [{H,' - AP,v,') - {H2' - AP2V2')]. (381) 

Heat Absorbed = Hi — H2. 
Steam initially wet 

Hi- H2= (xiri + qi) - {x2r2 + ^2). (382) 

X2 as calculated from equation (374). 

Steam initially dry, substitute o^i = 1. 

Steam initially superheated but wet at end of expansion 

Hi' - H2 = H,' - {X2r2 + q2). (383) 

Steam superheated throughout expansion, heat absorbed = 

Hi' - H2'. (384) 

Example 83. One pound of steam at initial pressure 115 pounds per 
square inch absolute and superheat 100 deg. fahr. expands adiabati- 
cally to 1 pound per square inch absolute. Required the various 
quantities at the final condition. 

From superheated steam tables for pi = 115 and ti = 438.1 = 
(338.1 + 100) we find: Hi' = 1243, Vi' = 4.51, iV/ = 1.6549. 

From saturated steam tables: P2 = 1, s = 333, ^2 = 69.8, H2 = 
1104.4, r2 = 1034.6, p2 = 972.9, 712 = 1.8427, 62 = 0.1327, 0-2 =0.016. 

N ' — 6 
• Final quality: X2 = ^ ^ 



712 

1.6549 - 0.1327 
1.8427 



0.826. 



i 



I 




TOTAL HEAT-ENTROPY DIAGRAM. 
g The ordinates are Total Heats. 
The abscissa are Entropies. 
.Vertical lines are lines of constant entropy. 
Horizontal lines are lines of constant total heat. 

I Reproduced by pennission from 

Masks and Davis' Stsam Tables. 



1.(36 



1.90 



1.94 



1.98 






iHBSRKSiSgS^SiiSSgiiSSi^iSlS^ 







ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 967 

Mollier diagram analysis, Fig. 622: Trace the intersection of pi = 115 
and ti = 438.1 vertically downward (constant entropy) to the line 
P2 = 1 and read 0.826 at the intersection of this line with the constant 
quality Hne (interpolated in this case). 

Final volume: V2 = X2U2 + o'2 

= 0.826 X 333 + 0.016 
= 275 cubic feet. 

(This quantity may be taken directly from the total heat pressure 
diagram.) 

External work: W = j [(^/ - APiVi') - {X2P2 + Q2)], 

1 4.4. V 11^ 

= 778 [(1243.1 - %C^^ 4.51) - (0.826 X 972.9 + 69.8)] 

77o 

= 213,938 foot pounds. 
Heat absorbed from the fluid 

= Hi - {X2r2 + 92) 

= 1243.1 - (0.826 X 1034.6 + 69.8) = 318.8 B.t.u. 

Mollier diagram, Fig. 622: Project the intersection of pi = 115 and 
ti = 438.1 upon the Y axis and read Hi = 1243. Similarly the pro- 
jection of the intersection of 7)2 = 1 and X2 = 0.826 gives H2 = 924.3, 
Hi' - H2= 1243 - 924.3 = 318.7 B.t.u. 

458. Polytropic Change of State. — A general law for the expansion of 
any vapor (wet, dry, or superheated) is 

pyn = constant, (385) 

gii^i" = V2V2'', (386) 



V2 



-"& <», 



By giving n special values we are able to obtain the various changes 
of state for constant volume, constant pressure, isothermal and 
adiabatic. 

The work done by expansion for all values of n, except n = 1, may be 
expressed 

W= rPdv^ (388) 

(389) 
Forn = 1, W = I Pdv (390) 

(391) 



= f PdV" 


P\Vi - P2V2 


n - 1 


= fpd. 


= Pit:, log. |. 



968 STEAM POWER PLANT ENGINEERING 

Saturated Steam. — Since with wet or saturated steam there can be 
no change of pressure without a change of temperature the value o^ 
n will vary with every change of state and for this reason the use of 
equations (385) and (388) are more troublesome than the preceding 
thermal analysis. An exception is that of ''saturated expansion" in 
which steam remains saturated throughout change of state. A study 
of the actual volume occupied by a pound of dry steam at various pres- 
sures will show that n has an approximately constant value of 1.0646 or, 
pi^,io646 = constant, (392) 

u = s — (T. (Except for high pressures the 
influence of a is negligible and 
u = s may be safely assumed.) 

This condition of constant saturation during expansion seldom occurs 
in steam engine practice but equation (392) offers the only simple 
solution of problems involving work done by such a change of state. 

Example 84. One pound of steam at an initial pressure of 115 
pounds per square inch absolute expands to a pressure of 2 pounds 
absolute and maintains a saturated condition throughout expansion. 
Required the final volume and the work done during expansion. 

From equations (386) and (392) 

1 



U2 



= (3.88 - 0.018) 



Work done W = 



1 

/115Y0646 

= 173.6 cubic feet. 

This value checks with that obtained 
from steam tables. 

PlUi — P2U2 

n - 1 
144 (115 X 3.862 - 2 X 173.6) 



1.0646 - 1 



= 216,000. 



Wet Steam. Actual Expansion. — The values of n for the expansion 
and compression curves of indicator diagrams from actual engines are 
subject to a wide variation. A study of several types and sizes of 
engines by J. Paul Clayton* gave values of n varying from 0.7 for wet 
steam to 1.34 for highly superheated steam. The average value of n 
is, however, not far from 1. That n = 1 for isothermal gas expansion 
and the average actual steam cylinder expansion is a mere coincidence 
and does not signify that the expansion in the latter is isothermal. 
See Conventional Diagram, par. 464. 

* University of Illinois Bulletin, Vol. 9, No. 26, 1915. 



ELEMENTARY THERMODYNAMICS — CHANGE OF STATE 969 

Example. — One pound of saturated steam at an initial pressure of 
115 pounds per square inch absolute expands so that its volume has 
been increased 5 times. Required the work done during expansion.* 

W = P,v, log. I 

= 144 X 115 X 3.88 log« |, 
= 103,200 foot pounds. 

Wet Steam. Adiahatic Expansion. — The ease with which prob- 
lems involving adiabatic expansion of vapor or moderately superheated 
steam can be solved by exact thermal analysis precludes the use of 
the more troublesome polytropic expansion law. A number of at- 
tempts have been made to derive laws which will give the value of n 
for adiabatic expansion of saturated or wet steam but their accuracy 
is limited to a comparatively narrow range of pressures and quality. 
A rule formulated by H. E. Stone f and often used in this connection is: 
n = 1.059 - 0.000315 p + (0.0706 + 0.000376 p) x. (393) 

Example 85. One pound of steam expands adiabatically from an 
initial pressure of 115 pounds per square inch and quality 0.9 to a 
pressure of 1 pound absolute. Required the final volume and the 
work done during expansion by exact thermal methods and by the 
polytropic law using equation (393) for determining the value of n. 

From steam tables : 

p, = 115, gi = 309, pi = 797.9, (9i = 0.4877, n^ = 1.103, Vi = 3.880, 
P2 = 1, 52 = 69.8, P2 = 972.9, 02 = 0.1327, ng = 1.9754, V2 = 333. 

Exact thermal methods: 

XiUi -^61 — 62 

Xi = 

ni 

_ 0.9 X 1.103 + 0.4877 - 0.1327 

1.8427 

= 0.785. 

V2 = X2U2 + (72 

= 0.785 (333 - 0.016) + 0.016 
= 261.4 cubic feet. 

^ = J [(^iPl + ^1) ~ (^2P2 + 92)] 

= 778 [(0.9 X 797.9 + 309) - (0.785 X 972.9 + 69.8)] 
= 149,843 foot pounds. 

Polytropic law: 

n = 1.059 - 0.000315 X 115 + (0.0706 + 0.000376 X 115) 0.9 

= 1.125. 
vi = XiUi + (71 = 0.9 X (3.88 - 0.016) + 0.016 
= 3.5 cubic feet. 

PlVl"" = P2V2''. 
115 X 3.51125 = 1 X V2^'^', 

V2 = 235.6 cubic feet. 

* Assuming n = 1. f University of Illinois Bulletin, Vol. 9, No. 26, p. 79. 



970 STEAM POWER PLANT ENGINEERING 

PiVi — P2V2 



W = 



n- 1 
144 (115 X 3.5 - 1 X 236.5) 



1.125 - 1 
= 181,232 foot pounds. 

The value of n which will give the same work during expansion 
according to the polytropic law as the exact thermal analysis for the 
conditions specified in the problem may be determined as follows: 

^^, PlVi - P2V2 

W = ^— j 1 

149,843 = 144(n5x3.5-lx261.4)_ 
n — 1 
n = 1.135. 

This value of n is an average only since the true value varies at 
different points along the expansion line. This may be shown by 
plotting the true adiabatic expansion line on logarithmic cross-section 
paper. See par. 465. 

Superheated Steam. Isothermal Expansion. — For steam so highly 
superheated that it does not approach the wet state at any point during 
the change of state, n = 1, and the exponential law offers the only 
simple solution for the work done during expansion. This case has 
been treated in par. 455. 

Superheated Steam. Adiabatic Expansion. — The work done during 
adiabatic expansion may be approximated from the polytropic law by 
making n = 1.3. Goodenough gives the following as more accurate 
than the simple law pV" = constant. 

p (v' + 0.088)1-31 = constant. (394) 

Example 86. Steam at 60 pounds per square inch absolute pressure 
and initially superheated to 300 deg. fahr. expands to a pressure of 
15 pounds absolute. Required the final volume and work done ac- 
cording to the polytropic law. 

From superheated steam tables for pi = 60 and superheat of 300 
deg. fahr. 

V = 10.41, 
60 (10.41 + 0.088)1-31 = 15 (v^' + 0.088)1^ 
V2' = 30.2. 

Thermal analysis gives V2 = 30. 

^ ^ Pi W + 0.088) + P2 W + 0.088) 
n — \ 
^ 144 (60 X 10.5 + 15 X 30.1) 

1.31 - 1 
= 83,000 approx. 

Thermal analysis gives W = 78,800, 



CHAPTER XXIV. — Supplementary 

ELEMENTARY THERMODYNAMICS OF THE STEAM ENGINE 

459. GeneraL — The recent marked improvement in the heat economy 
of piston engine is largely due to a better understanding of the ther- 
modynamic principles involved in its operation. Once constructed no 
amount of attention or mechanical adjustment will appreciably affect 
the economy since the heat efficiency is primarily a function of the 
design. It is not the object of this chapter to analyze the various 
thermodynamic laws underlying the design and operation of the piston 
engine but rather to show their application to the existing types of 
steam prime movers. In developing an engine with a view of better- 
ing the performance a knowledge is necessary of the theoretical limi- 
tations of the particular type under consideration. With this hmit as 
a guide the degree of perfection of the actual mechanism is readily 
ascertained by comparing test results with those theoretically obtain- 
able. Complete conversion of the heat supplied into useful work is 
impossible for even the perfect or ideal engine, hence some other 
standard than the heat supphed is desirable for comparison. There 
are several ideal cycles which simulate to a certain extent the action 
of steam in the real engine. The more important of these will be 
treated in detail. 

460. Carnot Cycle. — The Carnot cycle gives the highest possible 
efficiency for any type of heat and it would seem to be the most de- 
sirable cycle for the steam engine, but, as will be shown later, there 



Notations: 
A- JL. 



p = lb. per sq. in. abs. 



P = lb. persq. ft. abs. x = quality 

of wet steam, 
s = specific volume of dry steam, 

lb. per cu. ft. 
V = specific volume of vapor, lb. 

per cu. ft. 
a = specific volume of water, lb. 

per cu. ft. 
u = increase in volume during 

evaporation, cu. ft. 
t = deg. fahr. above zero. T = 

deg. fahr, abs. 
Cm = mean specific heat of water. 
C = mean specific heat of super- 
heated steam. 



H = heat content above 32 deg. fahr., 

B.t.u. per lb. 
X = total heat of dry steam, B.t.u. per 

lb. 
r = latent heat of vaporization, B.t.u. 

per lb. 
p = internal latent heat, B.t.u. per 

lb. 
q = heat of Hquid, B.t.u. per lb. 
6 = entropy of the liquid. 
n = entropy of the vapor. 
A^ = total entropy. 
Prime marks indicate superheat. 
Subscripts 1, 2, w, s indicate, respec- 
tively, initial condition, final condi- 
tion, wet steam, and superheated 
steam. 



971 




972 STEAM POWER PLANT ENGINEERING 

are practical limitations which more than offset the thermodynamic 
advantage. Nevertheless a study of this cycle is of importance in 
showing the absolute degree of perfection which can be realized 
theoretically. 

The diagram in Fig. 623 represents the pressure-volume action or 
indicator card of an ideal steam engine cylinder operating in the Carnot 
cycle. For simpUcity assume the cylinder to be one square foot in 
area, to contain unit weight of water and to have a piston displacement 
equivalent to one pound of saturated steam at the existing back pres- 
sure. At the beginning of the stroke the nonconducting cylinder 

contains water at temperature Ti 
corresponding to pressure Pi. Heat 
is added to the hquid until vapori- 
zation is complete, the movement 
of the frictionless piston being such 
that the pressure and therefore the 
temperature is constant, that is, ex- 
voiume ' pansion from to i is isothermal. 

Fig. 623. Indicator Card for Perfect ^^e source of heat is now removed 
Engine Operating in the Carnot Cycle. ^^^ ^^^ ^.^^^^ .^ ^^^^^^ ^^^^ ^ ^^ 

2 by the expansion of the steam. Since the cylinder is nonconduct- 
ing and there is no reception or rejection of heat the expansion from 1 
to 2 is adiabatic. From 2 to 3 heat is abstracted from the steam at 
such a rate that the temperature and hence the pressure remain con- 
stant, that is, the steam is compressed isothermally. At 3 the heat 
abstraction is terminated and the mixture of vapor and liquid is com- 
pressed adiabatically to the initial temperature and pressure Ti. The 
location of point 3 is such that water only at temperature Ti will be 
present at the end of compression. This assumption that there is only 
water at and saturated steam at 1 is not necessary and any degree 
of wetness or superheat may be assumed since it in no way affects 
the efficiency. 

The net work per cycle is represented by the shaded area 0123. 

Area 0123 = area 01 fd + area 12gf — area 32ge — area d03e (395) 
Area Olfd = PiVi = Pi {si — en) = PiUi. See equation (347). 

Since no heat is added during expansion from 1 to 2 the internal 
work is equal to the difference in intrinsic energy. See equation (379), 
hence: 

Area 12gf = [(pi + qO - {x,p, + gs)] j- (396) 

Area 32ge = P^vz - P2V2 = P^v^. (397) 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 973 

But V2 = X2U2 + 0-2 (see equation (398)) 

and ^4 = X3U2 + 0-2. 

Substituting these values in equation (397) 

Area 3^ge = P2X2U2 — P2X3U2 

= P2U2 {X2 - Xi). 

Since no heat is added during compression from 3 to and there is 
only liquid at the external work done on the steam is equal to the 
increase in intrinsic energy, or 

Area dOSe = [qi - (X3P2 + 52)] y- 

All of these factors with the exception of X2 and Xz may be obtained 
directly from the steam tables. X2 and Xz may be calculated from 
equation (374) or they may be taken directly from the temperature- 
entropy diagram. 

From the above data the PV diagram or indicator card may be 
readily plotted to scale. In order to obtain the true contour of the 
expansion and compression lines several intermediate points should 
be calculated and located on the diagram. 

The area 0123 when correctly drawn should check with the calcu- 
lated work. Substituting the values of the different areas in equation 
(395) we have 

Net work per cycle = PiUi + [(pi + gO - {X2P2 + ^2)] -r — P2U2 {x2 — x^) 

- [gi - {XzP2 + 72)] 2 

= Pi^i + J - ^2 {P2U2 + f )+ ^iP^U2 + ^ • (398) 

Heat absorbed in doing work 

= APiUi -\- pi - X2 {AP2U2 + P2) + xz (AP2U2 + P2), 
= APiUi + pi - (0:2 - X3) {AP2U2 + P2). (399) 

From equation (325) APiUi + pi = ri and AP2U2 + P2 = ^2. 
Therefore heat absorbed 

= n - r2 {xo - X3). (400) 

The water rate or steam consumption per hp-hr. of the ideal engine 
working in this cycle is 

W — Heat equivalent of 1 hp-hr. //inn 

Heat absorbed per lb. of fluid 

'''' (402) 



^1 - rs (0:2 - Xs) 



974 



STEAM POWER PLANT ENGINEERING 



Efficiency: 



E 



Heat absorbed 
Heat supplied 

n - rs (X2 - X3) 



Tl 



But 



Therefore 



T, 



r2 (x2 — xz) = jfr ^h see equation (368). 
-t 1 



E = 



n - Trrn 
^ 2 

ri 



7^1 



(403) 
(404) 
(405) 

(406) 




which is independent of the nature of the working substance and 
dependent only on the range of temperature. 

The shaded area 0123, Fig. 624, represents the indicator card of Fig. 
623 plotted in the temperature-entropy diagram in which ordinates 

are absolute temperatures and abscissas in- 
crease of entropy. This diagram is useful 
in visualizing the thermal changes per stroke 
or cycle. Line ww represents the increase 
of entropy of the liquid above 32 deg. fahr. 
and ss the increase of entropy of the vapor. 
Both of these lines are readily constructed 
by plotting several values of 6 and A^ as 
abscissas for corresponding values of T as 
ordinates. These quantities may be taken 
directly from steam tables. 0-1 therefore 
represents the isothermal expansion of the 
fluid from water at temperature T^ to dry 
steam at the same temperature. Since the entropy is constant for 
adiabatic expansion 1-2 represents the expansion of the saturated fluid 
from temperature Ti to temperature T2. Similarly 2-3 represents iso- 
thermal compression at temperature T2 and 3-0 adiabatic compression 
from temperature T2 to the initial condition. If the various lines are 
drawn to scale 

Heat supplied above 32 deg. fahr. = area mOln. 

Area mOln = 0-1 X Ti = UiTi = n. 

Heat rejected above 32 deg. fahr. = area m32n. 

Area m32n = 3-2 X T2 = riiTi. 

Heat absorbed = area 0123 = area mOln — area 1713211 
= n- niT2. 
= ri-T2 {X2 — X3). 



/ m 



Entropy 

Fig, 624. Temperature-en- 
tropy Diagram; Perfect 
Engine, Carnot Cycle. 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 975 
Quality at end of expansion x^ = — = 



ce ce ti2 

^ ,..,,. . - . c3 aO -he d, - $2 
Quality at beginning of compression x^ = — = = 

For any degree of wetness at the beginning and end of isothermal 
expansion the point will lie to the right of the intersection of ivw 
and Ti, and the point 1 will lie to the left of the intersection of ss and 
Ti. The figure 0123, however, will always be a rectangle. 

If isothermal apphcation of heat is continued during admission until 
the fluid is superheated the point 1 will still lie on the line aOl but to 
the right of the vapor Hne ss. In order to maintain a constant tem- 
perature of Ti in the superheated zone, the pressure must be lowered 
according to the law expressed by equation (308). Since superheat is 
supplied in practice with gradually increasing temperature and not 
isothermally the Carnot cycle is not a satisfactory standard for com- 
paring engines using superheated steam and hence this case will not 
be considered. 

Example 87. Determine the heat absorbed, water rate and effi- 
ciency of a perfect engine working in the Carnot cycle if the cyhnder 
contains only water at the beginning of the cycle and saturated steam 
at cut off. Initial pressure 215 lb. per sq. in. absolute; back pressure, 
2 lb. absolute. Assume one pound of fluid per cycle. 

From steam tables: 

pi = 215, ^1 = 388, si = 2.138, qi = 361.4, n = 837.9, pi = 754, 

01 = 0.5513, ni = 0.9885, en = 0.0185, Ni = 2.138, 

P2 = 2,t2= 126.15, S2 = 173.5, (?2 = 94, r^ = 1021, p2 = 956.7, 

02 = 0.1749, 712 = 1.7431, (72 = 0.0162. 

Qualities : 
Xo = zero. Xi = unity. 

X, = ^^1^' = ^-^^f ~3°'J^^^ = 0.7833. (See equation (374).) 
0,-e, 0..5513 - 0.1749 

Specific volumes: 

vo = (Ti = 0.0185. 

vi = si- ai = 2.138 - 0.0185 = 2.12. 

V2 = X2U2 + (72 = 0.7833 X 173.5 = 135.9. (See note, equation (310).) 

Vz = V2-v^= 135.9 - 37.53 = 98.37. 

V3 = XzU2 + (72 = 0.216 X 173.5 = 37.53. (See note, eauation r310).^ 



976 STEAM POWER PLANT ENGINEERING 

Work: 

Admission: Pi^i = 144 X 215 X 2.12 
= 65,625 ft. lb. 

Expansion = — [(pi + q^) — (x2P2 + ^2)] 

= 778 [(754 + 361.4) - (0.7833 X 956.7 + 94)] 
= 211,616 ft. lb. 
Exhaust: Ps^s = 144 X 2 X 98.37 
= 28,350 ft. lb. 

Compression = j [(gi — {X3P2 + ^2)] 

= 778 [361.4 - (0.216 X 956.7 + 94)], 
= 47,302 ft. lb. 
Net work = (65,635 + 211,616) - (28,350 + 47,302), 
= 201,599 ft. lb. 
Heat : 

Equivalent of work done = 201,599 -^ 778 = 259.1 B.t.u. 
Supplied = n = 837.8 B.t.u. 

Efficiency: Er = Hl^ = 0.309 = 30.9 per cent. 

2546 
Water rate: Wr = 7^777-^ = 9.83 lb. per hp-hr. 
259.1 

Temperature-Entropy Diagram. 

Heat equivalent of work done = ni{Ti — T^) = Ui {ti — ^2) 

= 0.9885 (388 - 126.15) 
= 259.0 B.t.u. 

Efficiency = ^' ~ ^' = ?^j^ = 0.309 = 30.9 per cent. 
1 1 o4o 

While it is conceivable to build an engine which will simulate the true 
Carnot cycle it would be practically impossible to do so without in- 
troducing evils which would more than counterbalance the thermo- 
dynamic advantage. The compression in the actual engine must not 
be confused with the adiabatic compression of the Carnot cycle since 
the cushion steam involved in the operation of the former is but a frac- 
tion of the total fed to the cylinder and has but little influence on the 
thermodynamic action of the engine. 

A modification of the Carnot cycle, known as the regenerative steam- 
engine cycle and which has the same efficiency as the former, has been 
simulated by a special type of Nordberg pumping engine. The engine 
is quadruple expansion with four cyhnders, three receivers and five 
feed-water heaters in series a, h, c, d, and e. The feed water is taken 
from the hot well and passed in succession through the various heaters: 
a receives its heat from the exhaust steam on its passage to the condenser; 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 977 




b receives its heat from the low-pressure cyHiuler jacket; aiul c, d, and 
e, respectively, from the third, second, and first receivers. Referring 
to Fig. 625, if 1-c' is drawn parallel to the water line ww the area Olc'c 
will equal the area of the Carnot cycle 0123. The Nordberg engine 
approximates this cycle as indicated 
by the broken lines. The expan- 
sion in the first stage corresponds 
to 1-ai, that in the second to ai-a2, 
and so on for each of the other 
stages. Heat represented by the 
area below ai-a/ is abstracted from 
the first stage and is used to raise 
the condition of the water from 62' 
to 61; heat corresponding to the 
area below a2«2' is withdrawn from 
the second stage and is used to 
raise the condition of the water 
from 63 to 62; and so on for each 
stage. Thus heat is abstracted by 
steps from the expanding steam 
and is used for progressively heat- 
ing the feed water. By increasing 
the number of steps the nearer will the actual cycle approach that of 
the ideal. The Nordberg compressor, Table 82, attained 73.7 per cent 
of the efficiency of the Carnot cycle for the same temperature limits and 
its heat economy has not yet been excelled. 

461. Rankine Cycle. Complete Expansion.* — This cycle has been 

adopted by the American So- 
ciety of Mechanical Engineers 
and the British Institution of 
Civil Engineers as the stand- 
ard for comparing the per- 
formance of all steam prime 
movers. It is of value not 
only in comparing the per- 
formances of steam engines 
with each other but also in 
comparing engines with tur- 
bines. In an engine working 
according to the Rankine cycle, steam is admitted at constant pressure, 

* This is often called the Clausius cycle since it was published simultaneously but 
independently by both Clausius and Rankine. 



Eatropy 
Fig. 625. Regenerative Steam Engine 
Cvcle. 




VolutLC 

Fig. 626. Indicator Card for Perfect Engine 
Working in the Rankine Cycle with Com- 
plete Expansion. ' 



978 STEAM POWER PLANT ENGINEERING 

expanded adiabatically to the back pressure and exhausted at that pres- 
sure. The engine has no clearance and there are no heat losses from 
friction, imperfect expansion, or otherwise, all the energy taken from 
the steam being converted into work. The diagram 0123, Fig. Q?"^ ^ 
represents the familiar indicator card or pressure-volume diagram of 
the working fluid operating in this cycle. 0-1 represents the admission 
of steam from the boilers at constant pressure Pi; 1-2 is an adiabatic 
expansion to exhaust pressure P2; 2-3 exhaust at constant pressure 
P2; and 3-0 a practically constant volume pressure rise. 
For all conditions of steam: 

Work done during admission = area Olfd 

Work done during expansion = area 12gf 

Work done during exhaust = area 32gd 

Net work = area Olfd — area 12gf — area 32 gd 
= area 0123 

Per pound of wet or saturated steam: 

Work done during admission = Pi {xiUi + ai) ft. lb. 

Work done during expansion = -j [{xipi + qi) — {X2P2 + 52)] ft. lb. 

Work done during exhaust = P2 (^2^2 + 0-2) ft. lb. 

Net work = Pi (xiUi + <^i) + T K^iPi + ^0 

- (^2P2 + 52)] - P2 {X2U2 + crs) ft. lb. (407) 
= xin -\-qi- fer2 + ^2)* B.t.u. (408) 
= Hi-H2 B.t.u. (409) 

Per pound of steam superheated at admission but wet or saturated 
at end of expansion: 

Work done during admission = PiVi ft. lb. 

Work done during expansion = {-tHi — PiVi] — -j {X2P2 + Q'2) ft. lb. 

Work done during exhaust = P2 {X2U2 + 0-2) ft. lb. 

Net work = P,v,' - TQ Hi' - P,v,'^ - (X2P2 + ^2)] 

— P2 (X2U2 + 0-2) ft. lb. 
= Hi — {X2P2 + ^2) — AP2ix2U2 + 0-2) B.t.u. 
= Hi'-{x2r2 + ^2)* B.t.u. (410) 

= ^/-i72 B.t.u. (411) 

* The quantities Pi<ri and P2<Ti found by reducing equation are negligible and have 
been omitted. 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 979 

Per pound of steam superheated throughout admission and expansion : 
Work done during admission = PiVi ft. lb. 

Work done during expansion = -jHi — PiVi — \-jIi2 — P2V'2.'\ii. lb. 
Work done during exhaust = P2V2 ft. lb. 

Net work = PW + ^ {H,' - H2') - Piv,' + P2V2' 



- P2V2' ft. lb. 
= H,' - H2' B.t.u. 



(412) 
(413) 



Calling Hi and Hn the initial and final heat content for all conditions 
of steam, a general expression for the heat converted into work Hw is 



Hw = Hi — Hn. 
Heat suppHed Ht above exhaust temperature t is 

Ht = Hi - qn. 



Efficiency Er = 



Hi — Hn 



Hi - Qn 

Steam consumption or water rate, lb. per hp-hr., is 

2546 



Wr- 



Hi — Hn 



(414) 

(415) 
(416) 

(417) 



The temperature-entropy diagrams for the conditions discussed above 
are shown in Figs. 627 to 629. For saturated or wet steam it will be 





H« 



Fig. 627. Temperature-entropy Dia- 
gram; Perfect Engine, Rankine 
Cycle with Complete Expansion. 
Steam Dry at Cut-off. 



Entropy 



Fig, 628. Temperature-entropy Dia- 
gram; Perfect Engine, Rankine Cycle 
for Wet Steam at Cut-off. 



noted that the admission hue is an isothermal since a constant pressure 
expansion for saturated steam is also a constant temperature one. 
For superheated steam, however, the temperature increases with the 



980 



STEAM POWER PLANT ENGINEERING 




Entropy 



degree of superheat, the pressure remaining constant, and the relation 
between pressure and volume varies according to the law expressed 
in equation (308), that is, the location of point 1', Fig. 629, is fixed by de- 
termining the entropy corresponding to pressure Pi and temperature Ti. 

This may be calculated from equa- 
tion (343) or it may be taken 
directly from superheated steam 
tables. 

A study of equation (416) in 
connection with the Mollier dia- 
gram will show that 

(1) The Rankine cycle when 
using superheated steam has a 
lower theoretical efficiency than 
that of the same cycle with satu- 
rated vapor having the same maxi- 
mum temperature. 

(2) The theoretical efficiency in- 
creases but slightly with the in- 

FiG. 629. Temperature-entropy Diagram: . u i. j.u 

„ „ ^ T^ • Ti , • ^ 1 r Ox crease m superheat, the maximum 
Perfect Engme, Rankme Cycle for Steam . . 

Superheated throughout Expansion. pressure remammg constant; see 

Table 76. 

(3) The theoretical efficiency increases rapidly with the increase in 
pressure range; see Table 71. 

The behavior of the actual engine under these conditions is discussed 
in paragraphs 179 and 182. 

A comparison of the Carnot and Rankine cycle shows a lower effi- 
ciency for the latter for the same operating conditions, as would be 
expected. The water rate for the Carnot cycle, however, is higher. 
This apparent anomaly is due to the fact that the heat supplied per 
pound of fluid is much larger in the Rankine than in the Carnot. 
Thus less weight of steam is used per hp-hr., but each pound receives 
more heat and this is used less efficiently. 

Example 88. A perfect engine operating in the Rankine cycle with 
complete expansion takes steam at 115 lb. per sq. in. absolute pressure, 
quality 98, and exhausts against a back pressure of 1 lb. absolute. 
Required the condition of the steam at end of expansion, the work 
done, efficiency, and water rate. 

From steam tables: 

pi = 115, ti = 338.1, n = 879.8, qi = 309, Hi = 1188.8, di = 0.4877, 

ni = 1.103, 
P2 = 1, t2 = 101.8, r2 = 1034.6, g2 = 69.8, 02 = 0.1327, 

712 = 1.8427, 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 981 



Xi 



X\n\ + ^1 



- (See equation (374).) 

ii'i 

0.98 X 1.103 + 0.4877 - 0.1327 



1.8427 



= 0.779, 



Heat converted into work 

= XxTx + gi - {xivi + ^2) 

= 0.98 X 879.8 + 309 - (0.779 X 1034.6 + 69.8) 

= 1171.2 - 875.7 = 295.5 B.t.u. per lb. 

Hi — Hi 



Efficiency 



Water rate = 



295.5 



1171.2 
2546 



69.8 



= 0.268 = 26.8 per cent. 



Hi — H2 
2546 



295.5 



= 8.62 lb. per hp-hr. 



The initial and final heat content may be taken directly from the 
MoUier diagram; as a matter of fact it is customary in practice to use 
the diagram except where extreme accuracy is necessary or when the 
given conditions are beyond the range of the charts. 

462. Rankine Cycle with Incomplete Expansion. — If expansion after 
cut-off is not carried far enough to reduce the pressure to that of the 
back pressure line as shown in Fig. 630 the Rankine cycle more nearly 
simulates the cycle of the actual 
engine. This cutting the ^'toe" 
off the diagram decreases the 
efficiency, but permits of the use 
of a smaller cyhnder. A com- 
parison of the diagram in Fig. 626 
with that in Fig. 630 will show 
that the area 012'3' is of the 
same outline as area 0123, conse- 
quently the work done would be 
that corresponding to complete 
expansion to pressure Pc plus 
that represented by area S'2'23. 

If He represents the heat content corresponding to complete expansion, 
to pressure Pc the heat equivalent of the work done (area 012'3') is 
Hi - He B.t.u. per lb. 

Work corresponding to area 3'2'23 = (Pc — P2) Ve ft. lb. per lb. 

Hence, heat converted into work = Hi — He -\- A {Pc — P2) Oc. 



Y 


.^— -— -1, 






i 


J 


m 


y^' 





id_ 


' i/ 


?= 


Jl 



Volume 

Fig. 630. Indicator Card for Perfect En- 
gine Working in the Rankine Cycle with 
Incomplete Expansion. 



982 



STEAM POWER PLANT ENGINEERING 



Heat supplied is the same as for complete expansion = Hi 

Hi - He + A(Pc- P,) vc 



Q2. 



Therefore efficiency E/ = 
Water rate W = 



Hi - qi 
2546 



(418) 



(419) 



Hi- He + A (P, - P2) vo 
For wet steam, Vc = XcU^ + 0-2 = XcS2 (for all practical purposes). 
For dry steam, ^c = S2 — 0-2. 
For superheated steam, Vc = V2' — a^. 

The temperature-entropy diagram differs from that for complete 
Y u, expansion in the curtailment of lines 1-3' and 

3' -3 by constant- volume pressure drop 2' -2, 
Fig. 631. 

Example 89. Same data and requirements 
as in preceding example except that release 
occurs at a pressure of 4 lb. absolute. 

From steam tables: pi and p2 as in preced- 
ing example, 

p, = 4, re = 1005.7, qo = 120.9, Be = 0.2198, 

7la = 1.6416, S2 = 90.5, 
XiTli -\- 01 — dc 



L 




\ 


1 


M 


1 


\ 




k\' 


^2 


■^ 


3' 


\ 


f ra 




? 


I i 


/ 



Eutropy 

Fig. 631. Temperature-en- 
tropy Diagram ; Perfect 
Engine, Rankine Cycle 
with Incomplete Expan- 
sion. Steam Dry at Cut- 
off. 



Xr, 



^ 0.98 X 1.103 -t- 0.4877 - 0.2198 
1.6416 

= XcS2 = 0.822 X 90.5 
= 74.4. 



He = XcTc + qc 

= 0.822 X 1005.7 + 120.9 
= 947.6. 
Hi = 1171.2 (same as in preceding example). 

Efficiency = H.-H. + AiP.-P.)v. 
Hi — q2 
1171.2 - 947.6 + }n (4 - 1) 74.4 



1171.2- 69. 
1171.2-947.6 + 41 



264.6 



1171.2 - 69.8 1101.4 

= 0.24 = 24 per cent. 

2546 
Water rate = ', ^ = 9.62 lb. per hp-hr. 
264.6 

463. Rankine Cycle with Rectangular PV-Diagram. — This cycle is 
the least efficient of all vapor cycles in practical use but represents the 
action of the fluid in direct-acting steam pumps, direct-acting air com- 
pressors and engines taking steam full stroke. It may be looked upon 
as a limiting case of the Rankine cycle. From Fig. 632 it is apparent 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 983 



that 

Work done = A (Pi - P2) v B.t.u. (420) 

For wet steam, v = XiUi -\- <ji = XiS\ (for most purposes). 
For dry steam, v = Si — o-i. y 

For superhetead steam, v = Vi' — ai. 

Heat received is the same as that in the „ 
Rankine cycle | 



Efficiency = 
Water rate = 



= Hi - q^. 

A (Pi - P2) V 



Hi — Qn 

2546 

A (Pi - P2) V 



(421) 
(422) 




Fig. 632. 



Example 90. A perfect direct-acting steam pump operating in the 
rectangular PV cycle takes steam at initial pressure 115 lb. per sq. in. 
absolute, quality 98 per cent and exhaust against a back pressure of 15 
lb. absolute. Required the work done per lb. of fluid, efficiency and 
the water rate. 

From steam tables: pi = 115, Si = 3.88, Hi = 1188.8, 
P2 = 15, Qn = q2 = 181.0. 
Heat converted into work = A (Pi — P2) XiSi 

= 7^1(115 - 15)0.98 X 3.88 



Efficiency 
Water rate 



70.4 B.t.u. 
70.4 



1188.8 - 180 

2546 _,. , , 

-^^^TTT = 36 lb. per hp-hr 



= 0.07 approx. = 7 per cent. 



464. Conventional Diagram. — In designing an engine it is customary 

to assume as a basis of reference an ideal cycle which considers only the 

Y , kinetic action of the steam 

in the cylinder. This per- 
mits of analysis without the 
use of steam tables. The 
expansion is assumed to 
be hyperbolic because the 
equilateral hyperbola is 
readily constructed and be- 
cause expansion in the 
actual engine conforms ap- 
proximately to the law PV" = C (see paragraph 458). According to 
the 1915 A.S.M.E. Code the ideal engine is assumed to have no clear- 
ance and no losses through wire-drawing during admission or release. 
The initial pressure is that of the boiler and the back pressure that of 
the atmosphere for a non-condensing engine, and of the condenser for 




984 STEAM POWER PLANT ENGINEERING 

a condensing engine. Such a diagram for a simple non-condensing en- 
gine is illustrated in Fig. 633. 0-1 represents admission at constant 
pressure Pi, 1-2 represents hyperbolic expansion from cut-off 1 to re- 
lease at 2 and 2-3 represents exhaust at atmospheric pressure P2. 
The work done is represented by the 

area 0123 = area 01 fd + area 12gf — area 32 gd, 
area 01 fd = PiVi, 

area 12gf = PiViloge — (see paragraph 458), 

area 32gd = P2V2. 

Therefore net work done 

W = P,v, (1 + loge ^j - P2V2, (423) 

letting 

— = r = ratio of expansion, 

W = PiVi (1 + loge r) - P2V2. (424) 

-_ ^ ^. „ ^resi 0123 
Mean effective pressure P^ = ' 

V2 

= y^ (1+log.r) -P2. (425) 

As the m.e.p. is generally used in pounds per square inch, dividing 
both members of the equation by 144 gives 

Pm = '^(l+\oger)-p2. (426) 

Theoretical maximum horsepower = J^ , (427) 

oo,UUU 

in which 

I = length of stroke, feet, 
a = area of cylinder, sq. in., 
n = number of working strokes. 

The ratio of the m.e.p. of the actual engine to that of the ideal dia- 
gram as determined above is called the diagram factor. This factor 
is determined by experiment and ranges as follows (Heat Power 
Engineering, Hirshfeld and Barnard, 1915, p. 325) : 

Simple slide-valve engine 55 to 90 per cent 

Simple Corliss engine 85 to 90 " " 

Compound slide-valve engine 55 to 80 " " 

Compound Corliss engine 75 to 85 " " 

Triple expansion engine 55 to 70 " . " 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 985 



The probable mean effective pressure for the engine under consider- 
ation is 

M.e.p. = ]),n X diagram factor. (428) 

Example 91. Determine the probable horsepower of a 12 inch X 12 
inch simple engine, 250 r.p.m., initial pressure 120 lb. per sq. in. absolute, 
cut off I stroke, diagram factor 0.75. 

Theoretical m.e.p. = i|^ (1 + log^ 4) - 15, 
= 56.53. 

56.53 X 0.75 = 42.4. 
42.4 X 1 X 113 X 500 



Probable actual m.e.p 
Probable i.hp. 



33,000 



^2.4. 



(2) 
(3) 
(4) 



465. Logarithmic Diagram. — It is a well-known fact that the equation 
of the polytropic curve PV" = C becomes a straight line when plotted 
on logarithmic cross-section paper and the slope of the line is the value 
of n. Conversely, when the expansion or compression curve of an' 
indicator becomes a straight line in the logarithmic diagram it shows 
that the change of state is in accordance with the law Py" = C. The 
logarithmic diagram derived from the indicator card is useful in analyzing 
cylinder performance and gives valuable information which cannot be 
readily obtained otherwise. Thus it has been demonstrated * that the 
logarithmic diagram is of great assistance in 
(1) Approximating clearance volume. 

Locating the stroke positions of cyclic events. 

Detecting leakage. 

Approximating steam consumption. 
Construction of the Logarithmic Diagram. — If the clearance volume 
is given the construction of the diagram is very simple. Draw the 
clearance line OY and the abso- ^ 
lute pressure line OX on the in- 
dicator diagram as illustrated in 
Fig. 634. Locate points i, ^, 5, etc. 
on the expansion line and tabulate 
the corresponding absolute pres- 
sures and volumes. For example, 
the pressure corresponding to point 
1 is Pi and its value is the length of 
the line Pi multiplied by the scale 
of the indicator spring. Similarly the volume corresponding to point 
1 is Vi and its value is the length of the line Vi multiplied by constant 

* A New Analysis of the Cylinder Performance of Reciprocating Engines. J. 
Paul Clayton, Univ. of 111. Bull. No. 26, Vol. 9, May 6, 1912. 





A 


B 


^ 


s^ 


v..4^ 


^^ 














\ 












D, 




^1 










Vd 










\'p' 


«,^ 








f-V^^ 


\ 






'' 






'^ 



Fig. 634. 



986 



STEAM POWER PLANT ENGINEERING 



m ( = piston displacement per stroke in cu. ft. divided by the length of 
the card I measured in inches). Transfer these points to logarithmic 
cross-section paper as illustrated in Fig. 635, using absolute pressures 
in lb. per sq. in. as ordinates and cu. ft. as abscissas. Repeat the 
operation for the compression curve and draw a smooth line through 

the various points. The ratio -7- (measured in inches) will be the value 

de 
of n for the expansion line and -Tf=n for compression. 

aj 





A 






B 


wH 




— 1 


y 






S 1 












-T^-'i 

_i \^ 


- 


n 


\ 








1 \ 






\ 








[> A 


c 




\^ 








\c 




i\ 












\ 


\ 


P 










\ 


\ 








' — 

e 




■N, 


_l__ ^ 


l 








D 

... 




X 



Volume -Cubic Feet 
Indicator Card — Logarithmic Diagram. 



Fig. 635. 



Approximating Clearance Volume. — If expansion and compression 
vary substantially according to the law PV" = C the clearance volume 
may be approximated by trial and error. All that is necessary is to 
assume different values of clearance and to plot the logarithmic diagram 
for each assumed value until the expansion or compression curve is a 
straight line. 

Locating the Stroke Position of Cyclic Events. — Except with a few 
types of four- valve engines it is difficult and oftentimes impossible to 
locate the points of cut-off, release, and compression from the indicator 
diagram. If there is no leakage the true points may be located on the 
logarithmic diagram by noting when the expansion and compression 
curves become straight; see Fig. 193, Chapter IX. 

Detecting Leakage. — The law Pv" = C is applicable only to cases 
where the weight of steam remains practically constant during change 
of state. When the weight changes materially as by leakage, the re- 
sulting expansion and compression hues on the logarithmic diagram 
depart from straight fines. This is clearly shown in Fig. 195. 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 987 




Fig. 636. 



Approximating Steam Consumption. — According to Clayton (1) 
there is a definite relation existing between Xc (quality at cut-off) and 
n in any one cylinder which is practically independent of cut-off posi- 
tion. (2) This relation is practically independent of cyhnder size and 
of engine speed; it is therefore 
applicable to other cylinders of 
the same type. (3) By means 
of the experimentally deter- 
mined relations of Xc and n, 
the value of Xc may be ap- 
proximated from the average 
value of n obtained from the 
expansion curves of one set of 
indicator diagrams taken simul- 
taneously; therefore the actual weight of steam present in one revo- 
lution may be approximated. (4) The actual steam consumption 
may be obtained by this method from the indicator diagram to within 
an average of 4 per cent of test measurements. These statements 
apply strictly to non-jacketed steam cylinders in good condition, ex- 
hausting at or near atmospheric pressure. In applying this method 
it is only necessary to determine 7i as previously outlined and find from 
the curve in Fig. 194 the corresponding value of Xc. Knowing the 
quahty of steam at cut-off the weight of fluid per stroke can be readily 
calculated. It will be noted that the curve in Fig. 194 is only an average 
approximation and that there is a considerable range in the values of 
Xc for a given value of n. By separating the points into groups of 
similar pressures and speeds, several lines coordinating n and Xc may be 
obtained and a greater accuracy is possible. For a complete discussion 
of this important subject consult Clayton's paper. 

466. Temperature-Entropy Diagram. — If the actual indicator card 
is transferred to the temperature-entropy chart the various heat ex- 
changes during expansion and compression may be seen at a glance. 
The area represented by the actual diagram, however, does not give 
the heat utilized in doing work since the weight of steam is not constant 
throughout the cycle. From cut-off to release the weight is constant 
if there is no leakage, as is the case from beginning of compression to 
admission, but the weights involved in each case are not the same. 
Therefore, only the expansion line shows the true behavior of all the 
steam used per cycle and the rest of the diagram is more or less conven- 
tional. The transfer of the pressure-volume to the temperature-en- 
tropy diagram is best illustrated by a specific example. 



988 STEAM POWER PLANT ENGINEERING 

Example 92. Curve 0123, Fig. 636, is an average indicator card taken 
from a 12 X 12 engine running at 300 r.p.m.; clearance volume 10 per 
cent; steam consumption by tests 2700 lb. per hr. Transfer the in- 
dicator card to temperature-entropy chart. 

Locate the zero clearance line OY and zero pressure lines OX, and 
measure the diagram as indicated. The cyUnder displacement per 

stroke = -^-r-. 7— = 0.785 cu. ft. 

144 X 4 

(1 2 + 4\ 
- — J — ^1 = 0.314 cu. ft. 

Weight of '^cushion steam" on the assumption that the steam is dry 
at the beginning of compression 

= 0.314 X 0.0498 = 0.0156 lb. 
(0.0498 = wt. of 1 cu. ft. of steam at 20 lb. abs. pressure.) 

Weight of steam used per stroke or ''cyhnder feed" 

2700 



600 X 60 



= 0.075 lb. 



Total weight of steam expanding = 0.075 -1- 0.0156 = 0.09 lb. 

Lay off saturation line mm. This line represents the volume of 0.09 
lb. of saturated steam for the various pressures within the range of the 
diagram. 

Draw several pressure lines such as abc and tabulate the ratio — . This 

ac 

ratio gives the quality of the steam at point b in the expansion curve 

f — = — = x). Note that — represents quahty only during expansion 

after cut-off and that it is simply a ratio for other parts of the cycle. 
^ Tabulate also the absolute temperature 

corresponding to the pressure under con- 
sideration. Next construct the water and 
c saturation curves ww and ss, respectively, 

as illustrated in Fig. 637. This may be 
-\ done conveniently by using absolute tem- 

^ peratures and entropies of water and vapor 

given in steam tables, the entropies being 
multiphed by 0.09. Locate point h' on 
the corresponding temperature line in such 

a position that rates —7- = — obtained 
X ct c ac 

YiQ 637 from the indicator card. The locus of the 

. point 6' will be the desired diagram. The 
thermal action during actual expansion is apparent from the diagram; 
thus it will be seen by inspection that the steam is wet at cut-off, 
that condensation takes place from 1 to 2' more rapidly than if expan- 
sion were adiabatic, and that reevaporation takes place from h' to 2. 

The foregoing analysis applies only to saturated or wet steam. In 
case of superheat the actual expansion will he beyond the saturation 



ELEMENTARY THERMODYNAMICS OF STEAM ENGINE 989 



curve as illustrated in Fig. 638 and the ratio — = — will not give the 

ac s 

quality. To find the temperature corresponding to v' multiply s, the 

specific volume of one pound of saturated steam at pressure P by the 

ratio — as measured from the diagram. From superheated steam 
ac 

tables or by means of equation (311) determine the temperature cor- 

v' 
responding to volume s X - and pressure P. To transfer the point b 

s 

to the temperature-entropy diagram 
draw the temperature line T" cor- 
responding to that just determined 
and locate point h' on this line such 
that cb^ = total entropy for pressure 





Fig. G38. 



Fk;. 639. 



P and temperature T\ The total entropy may be taken from super- 
heated steam tables or it may be calculated from equation (342). For 
the problem under consideration the entropy thus obtained must be 
multiplied by 0.09, the weight of fluid expanding per cycle. The locus 
of the point b' will be the desired diagram. 

466a. Steam Accounted for by Indicator Diagrams at Points near Cut- 
off and Release. — The steam accounted for, expressed in pounds per 
i.hp. per hour, may readily be found by using the equation 



^^^ \{C + E) Wc 
m.e.p. 



{H -h E) Wh], 



(429) 



in which 

m.e.p. = mean effective pressure, 

C = proportion of direct stroke completed at points on expan- 
sion line near cut-off or release, 
E = proportion of clearance, 

H = proportion of return stroke uncompleted at point on com- 
pression line just after exhaust closure, 
Wc = weight of 1 cu. ft. steam at pressure shown at cut-off or 

release point, 
Wh = weight of 1 cu. ft. steam at pressure shown at compression 
point. 



990 



STEAM POWER PLANT ENGINEERING 



The points near cut-off release and compression referred to are in- 
dicated in Fig. 640. 

In multiple expansion engines the mean effective pressure to be used 
in the above formula is the aggregate m.e.p. referred to the cylinder 
under consideration. In a compound engine the aggregate m.e.p. for 

the h-p. cylinder is the sum of the 
actual m.e.p. of the h-p. cylinder 
and that of the 1-p. cylinder multi- 
plied by the cylinder ratio. Like- 
wise the aggregate m.e.p. for the 1-p. 
cylinder is the sum of the actual 
m.e.p. of the 1-p. cyhnder and the 
m.e.p. of the h-p. cyhnder divided 
by the cylinder ratio. 
The relation between the weight of steam shown by the indicator 
at any point in the expansion hne and the weight of the mixture of 
steam and water in the cylinder may be represented graphically by 
plotting on the diagram a saturated steam curve showing the total 
consumption per stroke (including steam retained at compression) and 
comparing the abscissas of the curve with the abscissas of the expan- 
sion hne, both measured from the line of no clearance. 




Compression 



Atmospheric Line 

Fig. 640. Points where ''Steam Ac- 
counted for by Indicator" is Computed. 



CHAPTER XXV. - Supplementary 

PROPERTIES OF AIR. — DRY, SATURATED, AND PARTIALLY SATURATED 

467. GeneraL — Tables and charts giving the simultaneous physical 
and thermal properties of dry and saturated air for various temperatures 
are of great assistance in solving problems relative to the design and 
performance of evaporative surface condensers, water-cooling appa- 
ratus and air-conditioning devices. Table 169 gives the properties 
of dry and saturated air for various temperatures ranging from to 
212 deg. fahr. and Figs. 461 and 462 give a complete psychrometric 
chart for all conditions of dry, saturated, and partially saturated air 
within a temperature range of 20 to 350 deg. fahr. These charts are 
extremely useful in avoiding laborious calculations. 

468. Dry Air. — The physical and thermal properties of dry air as 
used in these tables and charts are based on the following laws estab- 
lished by the latest experiments with gases and vapors: 

^^" = constant = 0.755, (430) 

J- a 

Cpa = 0.2411 + 0.0000045 {h + ^), (431) 

Ha = Cpa (^2 - ^l), (432) 

in which 

Pa = absolute pressure of the dry air, in. of mercury, 
Va = volume of 1 lb. of dry air, cu. ft., 
Ta = absolute temperature of the air, deg. fahr., 
Cpa = mean specific heat of air at constant pressure between tem- 
peratures ti and t2, 
Ha = heat content, B.t.u. per 11). of air above temperature ^i, 
ti = initial temperature, deg. fahr., 
tz = final temperature, deg. fahr. 

A sample calculation of the properties of dry air as Hsted in Table 
169 is given in Example 93. 

Example 93. Required the specific volume and density of dry air 
at 100 deg fahr. under standard atmospheric pressure (= 29.92 in.). 
Required also the heat content per lb. above deg. fahr. 

991 



992 



STEAM POWER PLANT ENGINEERING 



TABLE 169. 

PROPERTIES OF SATURATED AIR. (Barometer 29.921.) 
Mixture of Air Saturated with Water Vapor. 



Tempera- 


Weight of 


Volume of 


Elastic 
Force of 
Vapor, In. 
of Mer- 
cury.* 


Elastic Force 
of the Dry Air 
in the Mixture, 
In. of Mer- 
cury. 


Weight of 1000 Cu. 


Ft., Lb. 


ture, De- 
grees 
Fahr. 


1000 Cu. Ft. 

of Dry Air, 

Pounds. 


One Lb. of 
Dry Air, 
Cu. Ft. 


Weight of 
the Dry 
Air, Con- 
tent. 


Weight of 
the Vapor, 
Content.* 


Total 
Weight of 
the Mix- 
ture. 


1 


2 


3 


4 


,5 


' 


7 


8 





86.35 


11.58 


0.037 


29.88 


86.23 


0.067 


86.90 


10 


84.53 


11.83 


0.063 


29.85 


84.31 


0.110 


84.42 


20 


82.71 


12.09 


0.103 


29.81 


82.44 


0.177 


82.62 


30 


81.04 


12.34 


0.165 


29.76 


80.62 


0.278 


80.90 


32 


80.71 


12.39 


0.181 


29.74 


80.24 


0.303 


80.54 


35 


80.19 


12.47 


0.203 


29.72 


79.70 


0.340 


80.04 


40 


79.43 


12.59 


0.248 


29.67 


78.77 


0.410 


79.18 


45 


78.61 


12.72 


0.300 


29.62 


77.86 


0.492 


78.35 


50 


77.88 


12.84 


0.362 


29.56 


76.94 


0.588 


77.53 


55 


77.10 


12.97 


0.436 


29.48 


75.98 


0.699 


76.68 


60 


76.33 


13.10 


0.521 


29.40 


75.05 


0.823 


75.88 


62 


76.04 


13.15 


0.560 


29.36 


74.66 


0.887 


75.54 


65 


75.64 


13.22 


0.622 


29.30 


74.08 


0.979 


75.06 


70 


74.91 


13.35 


0.739 


29.18 


73.08 


1.153 


74.23 


72 


74.63 


13.40 


0.790 


29.13 


72.68 


1.229 


73.90 


75 


74.24 


13.48 


0.874 


29.05 


72.08 


1.352 


73.42 


80 


73.53 


13.60 


1.031 


28.89 


71.01 


1.580 


72.59 


85 


72.83 


13.73 


1.212 


28.71 


69.92 


1.841 


71.76 


90 


72.15 


13.86 


1.421 


28.50 


68.78 


2.137 


70.92 


95 


71.53 


13.98 


1.659 


28.26 


67.59 


2.474 


70.06 


100 


70.87 


14.11 


1.931 


27.99 


66.34 


2.855 


69.19 


105 


70.22 


14.24 


2.241 


27.69 


65.05 


3.285 


68.33 


110 


69.64 


14.36 


2.594 


27.33 


63.64 


3.769 


67.41 


115 


69.01 


14.49 


2.993 


26.93 


62.16 


4.312 


66.47 


120 


68.40 


14.62 


3.444 


26.48 


60.60 . 


4.920 


65.52 


125 


67.80 


14.75 


3.952 


25.97 


58.92 


5.599 


64.52 


130 


67.20 


14.88 


4.523 


25.40 


57.14 


6.356 


63.50 


135 


66.67 


15.00 


5.163 


24.76 


55.23 


7.187 


62.43 


140 


66.09 


15.13 


5.878 


24.04 


53.18 


8.130 


61.31 


145 


65.53 


15.26 


6.677 


23.25 


51.01 


9.160 


60.17 . 


150 


64.98 


15.39 


7.566 


22.35 


48.63 


10.30 


58.93 


155 


64.43 


15.52 


8.554 


21.37 


46.12 


11.56 


57.68 


160 


63.94 


15.64 


9.649 


20.27 


43.39 


12.94 


56.33 


165 


63.41 


15.77 


10.86 


19.06 


40.47 


14.45 


54.92 


170 


62.89 


15.90 


12.20 


17.72 


37.33 


16.11 


53.44 


175 


62.46 


16.03 


13.67 


16.25 


33.96 


17.93 


51.89 


180 


61.88 


16.16 


15.29 


14.63 


30.34 


19.91 


50.25 


185 


61.42 


16.28 


17.07 


12.85 


26.44 


22.06 


48.50 


190 


60.94 


16.41 


19.01 


10.91 


22.26 


24.41 


46.67 


195 


60.61 


16.50 


21.14 


8.78 


17.17 


26.96 


44.13 


200 


59.98 


16.67 


23.46 


6.46 


12.97 


29.72 


42.69 


205 


59.74 


16.74 


^26.00 


3.92 


7.82 


32.71 


40.53 


210 


59.31 


16.86 


28.75 


1.17 


2.30 


35.94 


38.24 


212 


59.10 


16.92 


29.92 








37.32 


37.32 



* Goodenough. 



PROPERTIES OF AIR 



993 



TABLE 169. 



Continued. 



Te iipera- 

ture, De- 

grea^ Fahr. 


Weight of 
Water Neces- 
sary to Satu- 
rate 100 Lb. of 
Dry Air. 


Volume of One 
Pound of Dry Air 
+ Vapor to Satu- 
rate it, Cubic Feet. 


Heat Content 
per Pound of 
Dry Air, B.t.u. 


Latent Heat of 
Vapor in One 
Lb. of Dry Air 
Saturated with 
Vapor, B.t.u. 


Heat Content of 

One Lb. of Dry 

Air Saturated with 

Vapor, B.t.u. 





0.078 


11.59 


0.000 


0.964 


0.964 


10 


0.131 


11.86 


2.411 


1.608 


4.019 


20 


0.214 


12.13 


4.823 


2.623 


7.446 


30 


0.344 


12.41 


7.234 


4.195 


11.429 


32 


0.378 


12.47 


7.716 


4.058 


11.783 


35 


0.427 


12.55 


8.44 


4.57 


13.02 


40 


0.520 


12.70 


9.65 


5.56 


15.21 


45 


0.632 


12.85 


10.86 


6.73 


17.59 


50 


0.764 


13.00 


12.07 


8.12 


20.19 


55 


0.920 


13.1') 


13.28 


9.76 


23.04 


60 


1.105 


13.33 


14.48 


11.69 


26.18 


62 


1.188 


13.40 


14.97 


12,12 


26.84 


65 


1.323 


13.50 


15.69 


13.96 


29.65 


70 


1.578 


13.69 


16.90 


16.61 


33.51 


72 


1.692 


13.76 


17.38 


17.79 


35.17 


75 


1.877 


13.88 


18.11 


19.71 


37.81 


80 


2.226 


14.09 


19.32 


23.31 


42.64 


85 


2.634 


14.31 


20.53 


27.51 


48.04 


90 


3.109 


14.55 


21.74 


32.39 


54.13 


95 


3.662 


14.80 


22.95 


38.06 


61.01 


100 


4.305 


15.08 


24.16 


44.63 


68.79 


105 


5.05 


15.39 


25.37 


52.26 


77.63 


110 


5.93 


15.73 


26.58 


61.11 


87.69 


115 


6.94 


16.10 


27.79 


71.40 


99.10 


120 


8.13 


16.52 


29.00 


83.37 


112.37 


125 


9.53 


16.99 


30.21 


97.33 


127.54 


130 


11.14 


17.53 


31.42 


113 64 


145 06 


135 


13.05 


18.13 


32.63 


132.71 


165.34 


140 


15.32 


18.84 


33.85 


155.37 


189.22 


145 


18.00 


19.64 


35.06 


182.05 


217.10 


150 


21.22 


20.60 


36.27 


214.03 


250.30 


155 


25.11 


21.73 


37.48 


252.61 


290.10 


160 


29.87 


23.09 


38.69 


299 . 55 


338.20 


165 


35.77 


24.75 


39.91 


357.75 


397.70 


170 


43.24 


26.84 


41.12 


431.20 


472.30 


175 


52.90 


29.51 


42.33 


526.0 


568.30 


180 


65.77 


33.04 


43.55 


651.9 


695.50 


185 


83.59 


37.89 


44.76 


826.1 


870.90 


190 


109.80 
191.00 

229.50 
419.00 


45.00 
56.20 

77.24 


45.97 
47.20 

48.40 
49.62 
50.83 
51.39 






195 






200 






205 






210 








212 





















994 STEAM POWER PLANT ENGINEERING 



From equation (430), 

29.92 X 7, 



0.755, 



100 + 459.6 

Va = 14.11 cu. ft. per lb. 

Density = = 0.071 lb. per cu. ft. 

From equation (431), 

Cpa = 0.2411 + 0.0000045 (0 + 100) = 0.2416, 

and from equation (432), 

Ha = 0.2416 (100 - 0) = 24.16 B.t.u. per lb. 

469. Saturated Air. — Water, if placed in a vacuum chamber, will 
evaporate until the pressure in the chamber has reached that of vapor 
corresponding to the temperature of the water. If the water is intro- 
duced into a chamber containing dry air the evaporation will proceed 
precisely the same as in the vacuum until the pressure has risen by an 
amount corresponding to the vapor pressure for the temperature. In 
this case, according to Dalton's law (paragraph 226) each substance 
will exert the pressure it would if alone occupying the volume, and the 
final pressure will be the sum of that of the vapor and that of the air. 
Air is said to be saturated with moisture when it contains the saturated 
vapor of water. It might be better to say that the space is saturated 
since the presence of air has no effect on the vapor (the temperatures 
being the same) other than that the air retards the diffusion of water 
particles. Perfectly dry air does not exist in nature since evaporation 
of water from the earth's surface causes the atmosphere to be more or 
less diluted with vapor. 

The weight of saturated water vapor per cubic foot depends only 
on the temperature and not on the presence of air. 

The various properties for air completely saturated with water 
vapor may be calculated by means of equations (430) to (432), and 
Dalton's law which may be expressed 

. , . , Pa + Pv = P, (433) 

m which 

Pa = absolute pressure of the dry air in the mixture, inches of 

mercury, 
Pv = absolute pressure of saturated steam at the temperature of 

the mixture, in., 
P = total pressure, which for atmospheric conditions = 29.921. 

Therefore, ^^ ^ ^ _ ^^ ^^3^^ 

Pv may be taken directly from steam tables. 



PROPERTIES OF AIR 995 

From equation (430), 

V =Va = ^^J (435) 

in which 

Va = volume of 1 lb. of dry air (plus vapor to saturate) at pressure 

Pa and absolute temperature Ta, 
V = volume of vapor in 1 lb. of dry air when saturated, cu. ft. 

Evidently ^^a = tt' 

' a 

in which 

Wa = weight of dry air in 1 cu. ft. of saturated mixture. 

The weight, Wv, of vapor in 1 cu. ft. of saturated mixture is the 
density of saturated vapor at pressure P„ and temperature Ta- This 
may be taken directly from steam tables. 

Total weight of mixture per cu. ft. = Wa -{- Wv. 

The weight, Wv', of vapor necessary to saturate 1 lb. of dry air, 

wj = Vw. = VaW.. (436) 

Heat content H' , or total heat in a mixture of 1 lb. of dry air satu- 
rated with water vapor, measured above deg. fahr., and not including 
the heat of liquid, is 

H' = C^a + r,w/, (437) 

in which 

ta = temperature of the mixture, deg. fahr., 

Tv = latent heat of saturated vapor at temperature ta and pressure Pv. 

An application of these formulas to the calculation of the various 
quantities in Table 169 for a temperature of 100 deg. fahr. is given in 
Example 94. 

Example 94. Required the following properties of atmospheric air 
completely saturated with water vapor when the temperature of the 
mixture is 100 deg. fahr. : Elastic force or pressure of the vapor and of 
the dry air in the mixture, volume of 1 lb. of dry air plus vapor to 
saturate it, weight of dry air and vapor in 1000 cu. ft. of mixture, 
weight of water necessary to saturate 100 lb. of dry air, latent heat of 
the vapor content of 1 lb. of mixture and the heat content of 1 11). 
of dry air saturated with vapor. 

Pressure of vapor in the mixture : 

Pv = 1.931 in. (from steam tables). 
Pressure of dry air in the mixture : 

Pa = P — Pv 

" = 29.921 - 1.931 = 27.99 in. 



996 



STEAM POWER PLANT ENGINEERING 



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998 STEAM POWER PLANT ENGINEERING 

Volume of 1 lb. of dry air saturated with vapor: 
0.755 Ta 



V = 

= 15.08 cu. ft. 



P -P. 

0.755 (100 + 459.6) 



29.921 - 1.931 
Weight of dry air in 1000 cu. ft. of saturated mixture: 

^« = ^ = T^ = 0.06634 lb. per cu. ft. 

V a 15. (Jo 

1000 Wa = 1000 X 0.06634 = 66.34 lb. 
Weight of water vapor in 1000 cu. ft. of mixture: 

Wv = 0.002855 lb. per cu. ft. (from steam tables), 
1000 w, = 1000 X 0.002855 = 2.855 lb. 

Total weight of 1000 cu. ft. of mixture: 

= 66.34 + 2.855 = 69.19+ lb. 

Weight of vapor necessary to saturate 100 lb. of dry air: 

lOv' = VWv = VaWv 

= 15.08 X 0.002855 = 0.04305 lb. per lb. of dry air 
= 0.04305 X 100 = 4.305 lb. per 100 lb. of dry air. 

Total heat of the dry air content, above deg. fahr. : 

Ha = C^a (100 - 0) 

= 0.2416 X 100 = 24.16 B.t.u. per lb. 
Latent heat of the vapor content : 

r^w^ = 1036.6 X 0.04305 = 44.63 B.t.u. 
Total heat of 1 lb. of dry air saturated with vapor: 

Hi = Ha -\- VvWv, 

= 24.16 + 44.63 = 68.79 B.t.u. 

470. Partially Saturated Air. — As previously stated air is said to be 
saturated with moisture when it contains the saturated vapor of water. 
In this condition the weight of vapor per cu. ft. corresponds to the 
density of saturated steam at the temperature of the mixture. If the 
body of air contains only a fraction of the weight of vapor correspond- 
ing to saturation it is said to be partially saturated and the fraction 
is called the relative humidity. Partially saturated air in reality con- 
tains superheated vapor since the temperature of the mixture, which 
is also that of the vapor is higher than that of saturated vapor cor- 
responding to the actual pressure of the vapor in the mixture. The 
water vapor in the atmosphere is usually superheated. If a partially 
saturated mixture of air and water vapor is cooled at constant pressure 
the mixture tends to become more and more saturated until at a cer- 



PROPERTIES OF AIR 999 

tain temperature called the dew point, condensation begins to take 
place. The pressure of saturated vapor corresponding to the dew 
point is substantially the same as the partial pressure of the super- 
heated vapor in the original mixture. 

The relative humidity or degree of saturation is ordinarily deter- 
mined by an instrument called the psychrometer, which consists of two 
thermometers suitably mounted, the bulb of one thermometer being 
covered by a close-fitting wick, which is kept moist, and the other 
being exposed directly to the air. There are two types in general use, 
the ''stationary," and the ''sling." In the former the two thermom- 
eters are suitably mounted and hung in the shade, and in the latter 
they are whirled at a rate of about 200 r.p.m. The sling psychrometer 
gives more reliable results than the stationary device. The "aspi- 
ration" psychrometer is used when more accurate results are required. 

If the air is saturated, no evaporation takes place from the wet bulb 
and the two thermometers read alike, but if it is only partially satu- 
rated evaporation occurs and the readings of the wet bulb thermometer 
are lower than those of the dry. Experiment * has shown (1) that 
when an isolated body of water is permitted to evaporate freely in the 
air it assumes the true wet bulb temperature, (2) that the heat content 
of the air and vapor mixture is a constant for a given wet bulb tem- 
perature irrespective of the initial temperature and humidity, and (3) 
that the heat given up by the water and absorbed by the air-vapor 
mixture may be expressed 

Twiwu, — w) = Cpa {td — L) + wCps {td — tw), (438) 

in which 

Tw = latent heat of vaporization at wet bulb temperature, tw, B.t.u. 

per lb., 
Ww = weight of vapor in 1 lb. of dry air when saturated at wet bulb 

temperature, tw, lb., 
w = actual weight of vapor contained in 1 lb. of dry air at dry 
bulb temperature, td. 

Cpa and Cps = mean specific heats, respectively, of the dry air and 
vapor between temperatures t^ and td. 
Transposing equation (438) and reducing, 

_ fwWw Cpg {td tyj) ^ (A.'iQ\ 

fw ^ PS Kyd txo) 

For low pressures, 

Cps = 0.42 + 0.00005 (td - tw). 
* Willis H. Carrier, Trans. A.S.M.E., Vol. 33, I91I, p. 1014. 



1000 STEAM POWER PLANT ENGINEERING 

■ ' At the low pressures under consideration Dalton's law may be 
assumed to hold good for vapors, thus 

P' = hPd, (440) 

in which 

h = relative humidity at temperature ta, 
D' = actual density of the vapor at temperature td, lb. per cu. ft. 
Dd = density of saturated vapor at temperature td, lb. per cu. ft. 
P' = hPd = actual pressure of the vapor in the mixture at temper- 
ature td, 
Pd = pressure of saturated vapor at temperature td. 

Other notations as previously defined. 

By combining equations (438) to (441) and solving for h (omitting 
a number of neghgible factors), Carrier (Trans. A.S.M.E., Vol. 33, 1911, 
p. 1023) has deduced the following expression: 

^-[^-- 2800''-?3l ]A' (^^2)* 

in which 

Pw = pressure of saturated vapor at wet bulb temperature, in., 
P = barometric pressure, in., 

d = temperature difference between the wet and dry bulb ther- 
mometers. 

Other notations as previously defined. 

Since, according to statement (2), the heat content, H', of 1 lb. of 
dry air at temperature td with relative humidity h is the same as that, 
H, of 1 lb. of dry air at wet bulb temperature, tw, when completely 
saturated, then 

H' = H = r^w^ + Cpatn.- (443) 

For any atmospheric pressure Pi, other than P the relative humidity 
will be 

h=^, (444) 

in which 

hi = relative humidity at pressure Pi, 
h = relative humidity at pressure P. 

The values in Figs. 641 and 642 are based on the foregoing analysis. 
An application of the equations formulas is given in Example 95. 

* This expression is for use in connection with the aspiration psychrometer. For 
the shng psychrometer substitute (2755 — 1.28 t-u,) for (2800 - 1.3 tw). 



PROPERTIES OF AIR 1001 

Example 95. Determine the following quantities for partially satu- 
rated atmospheric air if the wet and dry bulb temperatures are 80 and 
100 deg. respectively: relative humidity, pressure of the vapor in the 
mixture, pressure of dry air and vapor content of the mixture, weight 
of 1000 cu. ft. of mixture, actual weight of vapor in 1 lb. of dry air, 
dew point, and the heat content of the mixture. 

Relative humidity: 

r (29.921 -1.029) 20 -1 _!_ 

^ L 2800- 1.3 X 80 J 1.931 

= 0.42 or 42 per cent. 

Vapor pressure in mixture: 

P' = hPd = 0.42 X 1.931 = 0.811 in. 
Dry air pressure in mixture: 

P^ = P - P' = 29.92 - 0.811 = 29.11 in. 
Volume of 1 lb. of dry air plus vapor content; see equation (435) : 

0.755 Ta 



F = F 

V y a p 

0.755 (100 + 459.6) 



= 14.5 cu. ft. 



29.11 
Weight of dry air in 1000 cu. ft. of mixture 

= 1000 4 = ^ = 68.96 lb. 
V 14.5 

Weight of water vapor in 1000 cu. ft. of mixture 

= 1000 X h X density of saturated steam at 100 deg. fahr. 
= 1000 X 0.42 X 0.00285 = 1.19 lb. 

Total weight of mixture 

= 68.96 + 1.19 = 70.15. 

Weight of vapor in 1 lb. of dry air: 
From equation (439) 

1047.4 X 0.02226 - 0.2416 X 20 ^ ^,^^ ,^ 

^ = 1047.4 + 0.429 X 20 = ^'^^^^ ^^' 

w may also be closely approximated as follows: 

w = hV X density of saturated vapor at 100 deg. fahr. 
= 0.42 X 14.5 X 0.002855 = 0.0174 lb. 

Total heat in 1 lb. of dry air containing w lb. of vapor at tempera- 
ture td'. 

From equation (443) 

H' = 1047.4 X 0.02226 + 0.2415 X 80 = 42.46 B.t.u. 

H' may also be approximated from the values in Table 169. 

H' = heat content of the dry air + /i X latent heat content of 
saturated vapor at temperature td = 100 
= 24.16 + 0.42 X 44.63 = 42.9 B.t.u. 



1002 STEAM POWER PLANT ENGINEERING 

An application of Table 169 and the psychrometric charts in Fig. 
641 is given in Examples 96 and 97. 

Example 96. Atmospheric air at 40 deg. fahr. and relative humidity 
0.80 is to be conditioned to 70 deg. fahr. and relative humidity 0.50. 
Determine the amount of moisture and heat to be added, (1) by means 
of Table 169 and (2) by means of the curves in Fig. 641. 

From Table 169: 

Original moisture content = 0.52 X 0.8 = 0.416 lb. per 100 lb. 

of dry air. 
Final moisture content = 1.578 X 0.5 = 0.789 lb. per 100 lb. of 

dry air. 
Moisture to be added = 0.789 - 0.416 = 0.373 lb. per 100 lb. 

of dry air. 

From Fig. 641: 

Initial moisture content (intersection of td = 40 and h = SO per 

cent) = 29 grains per lb. of dry air. 
Final moisture content (intersection of td = 70 and h = 50 per 

cent) = 55 grains per lb. of dry air. 

(^p; _ 2Q\ 
^ 1 = 0.371 lb. per 100 lb. 

of dry air. (7000 = grains per lb.) 

From Table 169: 

Initial heat content = 9.65 + 0.8 X 5.56 = 14.1 B.t.u. per lb. 
Final heat content = 16.90 + 0.5 X 13.96 = 23.88 B.t.u. per lb. 
Heat required = 23.88 - 14.1 = 19.78 B.t.u. per lb. of dry air. 

From Fig. 461 : 

Initial heat content (intersection of td = 40 and h = SO per cent) 
gives wet bulb t^ = 37.5; follow constant temperature line 
tw = 37.5 until it intersects saturation line h = 100 per cent; 
trace vertically upward to intersection of ''total heat" hne and 
read from marginal notation 14.1 B.t.u. per lb. 

Final heat content (intersection oi td = 70 and h = 40 per cent) 
gives tw = 55.8; follow constant temperature line /«, = 55.8 
until it intersects line h = 100 per cent; trace vertically up- 
ward to intersection of ''total heat" Une and read 23.88. 

The charts in Figs. 641 and 642 are reproduced to a greatly reduced 
scale and the readings cannot be made with the accuracy indicated 
in the example. In the original charts the wet and dry bulb tem- 
perature can be read to an accuracy of 0.1 degree and the other quan- 
tities proportionately. 

Example 97. Atmospheric air at 90 deg. fahr. and relative humidity 
of 80 per cent is to be conditioned to 70 deg. fahr. and 50 per cent 
relative humidity. Determine the temperature to which the original 
mixture must be reduced in order to have a relative humidity of 50 



PROPERTIES OF AIR 1003 

per cent when heated to 70 deg. fahr. Determine also the amount 
of heat to be abstracted to effect the initial cooling and tliat to be 
supplied to bring it to the final desired condition. 

Moisture content at td = 90 and /i = 0.8 = 3.109 X 0.8 = 2.487 lb. 
per 100 lb. of dry air, corresponding dew point = 83 deg. fahr., that is, 
at 83 deg. fahr. condensation begins. 

Moisture content at td = 70 and h = 0.5 = 1.578 X 0.5 = 0.789 lb. 
per 100 lb. of dry air. Corresponding dew point = 51.8 deg. fahr. 
This is the temperature to which the air must be cooled in order to 
have the required humidity when reheated to 70 deg. fahr. Heat 
content at td = 90 and h = 0.8 = 21.74 + 0.8 X 32.39 = 47.G5 B.t.u. 
per lb. 

Heat content at td = 51.8 and h = 1.0 = 21.19 B.t.u. per lb. 

Heat to be removed from water condensed due to cooling from 83 

. n o ^ f u 2.487 - 0.789 ^ 83 - 51.8 ^ .^ t. . ik 

to 51.8 deg. fahr. = :r^ X ^ = ^-^2 B.t.u. per lb. 

(This is comparatively small and may be omitted.) 

Total heat to be removed in cooKng from initial conditions to 51.8 
deg. fahr. = 47.65 - (21.19 + 0.42) = 26.04 B.t.u. per lb. 

Heat content at ta = 70 and h = 0.5 = 16.9 + 0.5 X 13.96 = 
23.88 B.t.u. 

Heat to be added to retemper from 51.8 to 70 deg. = 23.88 - 21.19 
= 2.69 B.t.u. per lb. 

These values, neglecting the heat of the liquid, may be taken directly 
from the curves in Fig. 641 as shown in the preceding example. 

Example 98. Evaporative Surface Condenser. — How many cubic 
feet of air and how many pounds of water spray must be forced through 
an evaporative surface condenser of the fan type in order to condense 
1000 pounds of steam per hour and maintain a vacuum of 25 inches, 
barometer 29? (Atmospheric air 80 deg. fahr., relative humidity 70 per 
cent.) The air and vapor issue from the discharge pipe under pressure 
of 4 inches of water, temperature 120 deg. fahr., relative humidity 
98 per cent. 

The absolute pressure in the condenser is 29.0 — 25.0 = 4 inches of 
mercury. 

The total heat to be withdrawn in order to cool and condense 1000 
pounds of steam per hour at absolute pressure of 4 inches to 120 deg. 
fahi*. is ^QQQ [1114.8 - (120 - 32)] = 1,026,000 B.t.u. 

Neglecting radiation and leakage losses, this is the heat to be ab- 
stracted per hour by the air and water spray. 
Air-vapor Mixture Entering Condenser. 
Pressure Pi of the dry air: 

Pi = 29.0 - 0.7 X 1.0314 = 28.28 in. 
(1.0314 = pressure of saturated vapor at temperature t = 
80. deg. fahr.) 

Volume Vi of 1 lb. of dry air plus its vapor content, equation (435) : 
F. = °:^^i(|g±M= 14.41 cu. ft. 



1004 STEAM POWER PLANT ENGINEERING ' 

Weight Wi of vapor in 1 lb. of dry air: 

wi = 0.7 X 14.41 X 0.00158 = 0.0159 lb. 
(0.00158 = density of saturated vapor at ^i = 80 deg. fahr.) 

Heat content Ha of 1 lb. of dry air above deg. fahr. : 

Ha = Cpah =• 0.2414 X 80 = 19.32 B.t.u. 

Latent heat Vy of vapor content in 1 lb. of dry air: 

r, = 0.7 (14.41 X 0.00158 X 1047.4) = 16.68 B.t.u. 
(1047.4 = latent heat of saturated vapor at temperature t = 80.) 

Total heat Hi of mixture in 1 lb. of dry air: 

Hi = 19.32 +16.68 = 36.00 B.t.u. 

Air-Vapor Mixture Leaving Condenser. 
Pressure P2 of the dry air : 

P2 = (29.0 + 0.294) - 0.98 X 3.444 = 25.92. 
(0.294 = value in inches of mercury of 4 inches of water pressure.) 

Volume V2 of 1 lb. of dry air plus its vapor content: 

^. 0.755 (459.6 + 120) ^„^„ „^ 

Weight W2 of vapor in 1 lb. of dry air: 

W2 = 0.98 X 16.89 X 0.00492 = 0.08143.] 
Heat content Ha of the dry air in 1 lb. of mixture: 

Ha' = Cpt = 0.2416 X 120 = 29.00 B.t.u. 
Latent heat r^' of vapor content in 1 lb. of dry air: 

r/ = 10.98 (16.89 X 0.00492 X 1025.6) = 83.53 B.t.u. 
Total heat H2 of the mixture in 1 lb. of dry air: 

H2 = 29.00 + 83.53 = 112.53 B.t.u. 

Heat taken up by 1 lb. of air plus water vapor in passing through the 
condenser 

= H2- Hi = 112.53 - 36.00 = 76.53 B.t.u. 

Total weight of dry air passing through condenser 

1,026,000 .o.AAlK u 

= rjr, ro = 13,400 lb. per hour. 

Total volume of air-vapor entering the condenser 

= 13,400 X 14.41 = 192,960 cu. ft. 
Water absorbed per lb. of dry air 

= W2-Wi = 0.08143 - 0.0159 = 0.06553 lb. 
Total moisture absorbed or weight of spray to be injected 

= 13,400 X 0.06553 = 878.0 lb. per hr. 



PROPERTIES OF AIR 1005 

For purpose of design it is sufficiently accurate to disregard the actual 
barometric pressure and assume it to be 29.92 inches. With this as- 
sumption the problem may be readily solved by means of Table 169 
or the curves in Figs. 461-2. 

From Fig. 461 (for U = 80 and h = 0.70) : 

Wet bulb = 72.2, Dew point = 69.0. 
wi= 107 grains = 0.0153 lb. 
Hi = 35.5 B.t.u. 
From Fig. 462 (for ^2 = 120 and hz = 0.98) : 

Wet bulb = 119.4, Dew point = 119.2, 
W2 = 555 grains = 0.0793 lb. 
H', = HI B.t.u. 

Moisture absorbed per lb. of dry air, and its vapor content, 

1^2 - w;i = 0.0793 - 0.0153 = 0.064 lb. 

Heat absorbed per lb. of dry air, and its vapor content, 

H^- Hi = 111 - 35.5 = 75.5 B.t.u. 

Since the moisture content per lb. of dry air at dew point is the same 
as that for all conditions of wet and dry bulb temperatures having 
that dew point temperature. 
From Table 169: 

wi for dew point 69.0 = 0.0152 lb. 
W2 for dew point 119.2 = 0.0793 lb. 

Moisture absorbed per lb. of dry air = 0.0793 - 0.0152 = 0.06411b. 
Since the heat content or total heat is constant for a given wet bulb 
temperature 

Hi for wet bulb 72.2 = 35.3 B.t.u. 
Ho for wet bulb 119.2 = 110.5. 

Heat absorbed per lb. of dry air and its vapor content 
H2- H = 110.5 - 35.3 = 75.2 B.t.u. per lb. 

These results check substantially with the calculated data. 

Example 99. Determine the quantity of air passing through the 
cooling tower and the weight of circulating water lost by evaporation in 
a surface-condensing power plant operating under the following con- 
ditions: Turbines, average load 1000 kw.; average water rate 20 lb. 
per kw-hr.; initial steam pressure 150 lb. abs.; superheat 50 deg. fahr.; 
vacuum 26.92 in.; barometer 29.92 in.; temperature of injection water, 
discharge water and outside air, 70, 100, and 65 deg. fahr., respectively; 
temperature of air leaving tower 90 deg. fahr.; wet bulb temperature 
of outside air and air leaving cooling tower 57 and 89 deg. fahr. respec- 
tively. 

Total heat to be abstracted from the steam = 

1000 X 20 ^223 - ?|^ - 105 * + 32^ = 19,580,000 B.t.u. per hr. 

* Assumed hot well temperature. 



1006 STEAM POWER PLANT ENGINEERINa 

Atmospheric air entering tower: 

From the curves in Fig. 461 (dry bulb temperature 65 deg. fahr. 

and wet bulb temperature 57 deg. fahr.). 
Moisture content of 1 lb. of dry air, Wi = 56 grains. 
Total heat of 1 lb. of dry air, with its vapor content, 

Hi = 24.3 B.t.u. 
Air- vapor mixture leaving tower : 

From the curves in Fig. 461 (dry bulb 90 and wet bulb 98). 
Moisture content of 1 lb. of dry air, W2 = 209 grains. 
Total heat of 1 lb. of dry air, with its vapor content, 

H2 = 52.8 B.t.u. 

Moisture absorbed by 1 lb. of dry air in passing through the tower 

= W2-W = 209 - 56 = 153 grains or 0.02186 lb. 

Heat absorbed by 1 lb. of dry air (plus its initial vapor content) in 
passing through the tower 

= H2 - H = 52.8 - 24.3 = 2^.5 B.t.u. 

Total weight of dry air required to abstract the heat from the circu- 
lating water 

1 9,580,000 .„_ „ „ „ ,, , 

= — "ooi — = 687,000 lb. per hr. 

Volume of 1 lb. of dry air and its vapor content entering tower 

„_. /459^6 + 65\ ,^_ ,, 
= 0.755 — 7^7^^^ — = 13.39 cu. ft. 
V 29.54 / 

(29.54 = pressure of the dry air in the mixture = 29.92 - 0.61 X 0.6218; 
0.61 = relative humidity and 0.6218 = pressure of saturated vapor at 
65 deg. fahr.) 

Total volume of atmospheric air entering tower 

687,000 X 13.39 ..onnn u 
= — - — WF. = 153,000 cu. ft. per mm. 



APPENDIX A 

DATA AND RESULTS OF EVAPORATIVE TEST 

A.S.M.E. Code of 1915 

(1) Test of boiler located at 

To determine 

Test conducted by 



Dimensions. 

(2) Number and kind of boilers 

(3) Kind of furnace 

(4) Grate surface (width length ) * sq. ft. 

(a) Approximatewidth of air openings in grate in. 

(6) Percentage of area of air openings to grate surface per cent 

(5) Water heating surface sq. ft. 

(6) Superheating surface sq. ft. 

(7) Total heating surface sq. ft. 

(a) Ratio of water heating surface to grate surface ( — ) to 1 

(6) Ratio of total heating surface to grate surface ( — ) to 1 

(c) Ratio of minimum draft area to grate surface 1 to ( — ) 

(d) Volume of combustion space between grate and heating surface, cu. ft. 

(e) Distance from center of grate to nearest heating surface ft. 

Date, Duration, etc. 

(8) Date 

(9) Duration hr. 

(10) Kind and size of coal 

Average Pressures, Temperatures, etc. 

(11) Steam pressure by gage lb. per sq. in. 

(a) Barometric pressure in. of mercury 

(12) Temperature of steam, if superheated deg. 

(a) Normal temperature of saturated steam deg. 

(13) Temperature of feed water entering boiler deg. 

(a) Temperature of feed water entering economizer deg. 

(6) Increase of temperature of water due to economizer deg. 

* Unless otherwise designated this is the total area enclosed within the furnace 
walls projected horizontally. 

1007 



1008 STEAM POWER PLANT ENGINEERING 

(14) Temperature of escaping gases leaving boiler deg. 

(a) Temperature of gases leaving economizer deg. 

(b) Decrease of temperature of gases due to economizer deg. 

(c) Temperature of furnace deg. 

(15) Force of draft between damper and boiler in. of water 

(a) Draft in main flue near boiler in. of water 

(6) Draft in main flue between economizer and chimney in. of water 

(c) Draft in furnace in. of water 

(d) Draft or blast in ash pit in. of water 

(16) State of weather 

(a) Temperature of external air deg. 

(6) Temperature of air entering ash pit * deg. 

(c) Relative humidity of air entering ash pit per cent 

Quality of Steam. 

(17) Percentage of moisture in steam or number of degrees of 

superheating per cent or deg. 

(18) Factor of correction for quality of steam 

Total Quantities. 

(19) Total weight of coal as fired t lb. 

(20) Percentage of moisture in coal as fired per cent 

(21) Total weight of dry coal (item 19 X [ ^ ~ iqq"^ ^^ 1) ^^' 

(22) Ash, clinkers, and refuse (dry) 

(A) Withdrawn from furnace and ash pit lb. 

{B) Withdrawn from tubes, flues, and combustion chamber lb. 

(C) Blown away with gases lb. 

(Z)) Total .lb. 

(a) Weight of clinkers contained in total ash lb. 

(23) Total combustible burned (Item 21 - Item 22D) J lb. 

(24) Percentage of ash and refuse based on dry coal per cent 

(25) Total weight of water fed to boiler § lb. 

(26) Total water evaporated, corrected for quality of steam (Item 25 

X Item 18) lb. 

(27) Factor of evaporation based on temperature of water entering boiler. . . . 

(28) Total equivalent evaporation from and at 212 deg. (Item 26 X 

Item 27) lb. 

* Thermometer should be protected from direct radiation of boiler and furnace. 

fThe term "as fired" means actual condition including moisture, corrected for 
estimated difference in weight of coal on the grate at beginning and end. 

t If either of the two items 22B and 22C is omitted, the fact should be so stated. 

§ Corrected for inequality of water level and of steam pressure at beginning and 
end. 



APPENDIX A 1009 

Hourly Quantities and Rates. 

(29) Dry coal per hour lb. 

(30) Dry coal per sq. ft. of grate surface per hour lb. 

(31) Water evaporated per hour, corrected for quality of steam lb. 

(32) Equivalent evaporation per hour from and at 212 deg.* lb. 

(33) Equivalent evaporation per hour from and at 212 deg. per sq. ft. 

of water heating surface * lb. 

Capacity. 

(34) Evaporation per hour from and at 212 deg. (same as Item 32) lb. 

(a) Boiler horsepower developed (Item 34 -r- 34^) b.hp. 

(35) Rated capacity per hour, from and at 212 deg lb. 

(a) Rated boiler horsepower b.hp. 

(36) Percentage of rated capacity developed per cent 

Economy. 

(37) Water fed per lb. of coal as' fired (Item 25 -^ Item 19) lb. 

(38) Water evaporated per lb. of dry coal (Item 26 -^ Item 21) lb. 

(39) Equivalent evaporation from and at 212 deg. per lb. of coal as fired 

(Item 28 H- Item 19) lb. 

(40) Equivalent evaporation from and at 212 deg. per lb. of dry coal 

(Item 28 -j- Item 21) lb. 

(41) Equivalent evaporation from and at 212 deg. per lb. of combustible 

(Item 28 -i- Item 23) lb. 

Efficiency. 

(42) Calorific value of 1 lb. of dry coal by calorimeter f B.t.u. 

(a) Calorific value of 1 lb. dry coal by analysis B.t.u. 

(43) Calorific value of 1 lb. of combustible by calorimeter B.t.u. 

(a) Calorific value of 1 lb. combustible by analysis B.t.u. 

(44) Efficiency of boiler, furnace, and grate 

r^^ ^ Item 40 X 970.4-1 

P^^ Item 42 J ^'' ''""^ 

(45) EflSciency based on combustible 

r^^ ^ Item 41 X 970.4-1 

L^Q^^ Item 43 J ^' ''""^ 

* The symbol "U. E., " meaning Units of Evaporation, may be substituted for the 
expression " Equivalent evaporation from and at 212° ". 

t If the calorific value is desired per lb. of coal ''as fired," multiply Item 42 by 

100 - Item 20 
100 



1010 STEAM POWER PLANT ENGINEERING 

Cost of Evaporation. 

(46) Cost of coal per ton of lb. delivered in boiler room dollars 

(47) Cost of coal required for evaporating 1000 lb. of water under ob- 

served conditions dollars 

(48) Cost of coal required for evaporating 1000 lb. of water from and 

at 212 deg. dollars 

Smoke Data. 

(49) Percentage of smoke as observed per cent 

(a) Weight of soot per hour obtained from smoke meter per cent 

Firing Data. 

(50) Kind of firing, whether spreading, alternate, or coking 

(a) Average thickness of fire in. 

(6) Average intervals between firings for each furnace during time 

when fires are in normal condition min. 

(c) Average interval between times of leveling or breaking up min. 

(51) Analysis of dry gases by volume 

(a) Carbon dioxide (CO2) per cent 

(6) Oxygen (O) per cent 

(c) Carbon monoxide (CO) per cent 

{d) Hydrogen and hydrocarbons per cent 

(e) Nitrogen, by difference (N) per cent 

(52) Proximate analysis of coal 

As fired. Dry coal. Combustible. 

(a) Moisture 

(&) Volatile matter 

(c) Fixed carbon 

(d) Ash 



100 per cent 100 per cent 100 per cent 
(e) Sulphur, separately determined referred to dry coal per cent 

(53) Ultimate analysis of dry coal 

(a) Carbon (C) per cent 

(6) Hydrogen (H) per cent 

(c) Oxygen (O) per cent 

(d) Nitrogen (N) per cent 

(e) Sulphur (S) per cent 

(f) Ash per cent 



(54) Analysis of ash and refuse, etc. ^^^ ^^" 

(a) Volatile matter per cent 

(6) Carbon per cent 

(c) Earthy matter per cent 



100 per cent 

{d) Sulphur, separately determined per cent 

(e) Fusing temperature of ash deg. 



APPENDIX A 

(55) Heat balance, based on dry coal 



1011 



(a) Heat absorbed by the boiler (Item 40 X 970.4) 

(6) Loss due to evaporation of moisture in coal 

(c) Loss due to heat carried away by steam formed by the 

burning of hydrogen 

(d) Loss due to heat carried away in the dry flue gases 

(e) Loss due to carbon monoxide 

(/) Loss due to combustible in ash and refuse 

(g) Loss due to heating moisture in air 

(h) Loss due to unconsumed hydrogen and hydrocarbons, to 

radiation, and unaccounted for 

(i) Total calorific value of 1 lb. of dry coal (Item 42) 



If it is desired that the heat balance be based on coal "as fired" or on "com- 
bustible burned" the items in the first column are multiplied by the proportion 




100 - Item 20 



100 
for "combustible burned." 



for coal "as fired" or by the proportion 



100 - Item 20 



100-(Item20+Item24) 



(1 
(2 
(3 
(4 
(5 
(6 
(7 
(8 

(9: 

(10 

(11 
(12; 

(13 

(14: 
(15; 
(16; 
(17; 
(18; 
(19; 

(20 
(21 



PRINCIPAL DATA AND RESULTS OF BOILER TEST. 

Grate surface (width length ) sq. ft. 

Total heating surface sq. ft. 

Date 

Duration hr. 

Kind and size of coal 

Steam pressure by gage lb. per sq. in. 

Temperature of feed water entering boiler deg. 

Percentage of moisture in steam or number of degrees of 

superheating per cent or deg. 

Percentage of moisture in coal per cent 

Dry coal per hour lb. 

Dry coal per sq. ft. of grate surface per hour lb. 

Equivalent evaporation per hour from and at 212 deg lb. 

Equivalent evaporation per hour from and at 212 deg. per sq. ft. 

of heating surface lb. 

Rated capacity per hour from and at 212 deg lb. 

Percentage of rated capacity developed per cent 

Equivalent evaporation from and at 212 deg. per lb. of dry coal lb. 

Equivalent evaporation from and at 212 deg. per lb. of combustible. . .lb. 

Calorific value of 1 lb. of dry coal by calorimeter B.t.u. 

Calorific value of 1 lb. of combustible by calorimeter B.t.u. 

Efficiency of boiler, furnace, and grate per cent 

Efficiency based on combustible per cent 



APPENDIX B 

DATA AND RESULTS OF STEAM-ENGINE TEST 

A.S.M.E. Code of 1915 



(1) Test of engine located at, 

To determine 

Test conducted by 



Dimensions, etc. 

(2) Type of engine (simple or multiple expansion) 

(3) Class of service (mill, marine, electric, etc.) 

(4) Auxiliaries (steam or electric driven) 

(a) Type and make of condenser equipment 

(6) Rated capacity of condenser equipment hp. 

(c) Type of oil pump, jacket pump, and reheater pump (direct or inde- 
pendently driven) : 

(5) Rated power of engine 

(o) Name of builders 

(6) Kind of valves 

(c) Type of governor 

1st 2d 3d 

(6) Diameter of cylinders in. 

(7) Stroke of pistons ft. 

(a) Diameter of piston-rod, each end . . in. 

(8) Clearance (average) in per cent of piston 

displacement 

(9) Hp. constant 1 lb. 1 rev hp. 

(a) CyUnder ratio (based on net piston 

displacement) 1 to" — 

(6) Area of interior steam surface, .sq. ft. 

(c) Area of jacketed surfaces sq. ft. 

(10) Capacity of generator or other apparatus 

consuming power of engine hp. 

Date and Duration. 

(11) Date 

(12) Duration hr. 

* For other matters relating to the analysis of engine performance, see treatises on 
thermodynamics . 

1012 



APPENDIX B 1013 

Average Pressures and Temperatures. 

(13) Pressure in steam pipe near throttle, by gage lb. per sq. in. 

(14) Barometric pressure in. of mercury 

(a) Pressure at boiler, by gage lb. per sq. in. 

(15) Pressure in 1st receiver, by gage lb. per sq. in. 

(16) Pressure in 2d receiver, by gage lb. per sq. in. 

(17) Pressure in exhaust pipe near engine by gage lb. per sq. in. 

(18) Vacuum in condenser in. of mercury 

(a) Corresponding absolute pressure lb. per sq. in. 

(19) Pressure in jackets and reheaters lb. per sq. in. 

(20) Temperature of steam near throttle deg. 

(a) Temperature of saturated steam at throttle pressure deg. 

(6) Temperature of steam leaving 1st receiver, if superheated deg. 

(c) Temperature of steam leaving 2d receiver, if superheated deg. 

(21) Temperature of steam in exhaust pipe near engine deg. 

(a) Temperature of injection or circulating water entering condenser . deg. 

(6) Temperature of injection leaving condenser deg. 

(c) Temperature of air in engine room deg. 

Quality of Steam. 

(22) Percentage of moisture in steam near throttle or number 

of degrees of superheating per cent or deg. 

Total Quantities. 

(23) Total water fed to boilers lb. 

(24) Total condensed steam from surface condenser (corrected for condenser 

leakage) lb. 

(25) Total dry steam consumed (Item 23 or 24 less moisture in steam) lb. 

Hourly Quantities. 

(26) Total water fed to boilers or drawn from surface condenser per hour . .lb. 

(27) Total dry steam consumed for all purposes per hour (Item 25 -i- 

Iteml2) lb. 

(28) Steam consumed per hour for all purposes foreign to the main engine, .lb. 

(29) Dry steam consumed by engine per hour (Item 27 — Item 28) lb. 

(a) Circulating water supplied to condenser per hour lb. 

Hourly Heat Data. 

(30) Heat units consumed by engine per hour [Item 29 X (total heat of 

steam per pound at pressure of Item 13 minus heat in 1 lb. of water 

at temperature of Item 21)] B.t.u. 



1014 



STEAM POWER PLANT ENGINEERING 



(a) Heat converted into work per hour B.t.u. 

(6) Heat rejected to condenser per hour (Item 29a X [Item 216 — 

21a]) (approximate) B.t.u. 

(c) Heat rejected in form of uncondensed steam withdrawn from 

cyHnders * B.t.u. 

(d) Heat lost by radiation B.t.u. 



Indicator Diagrams. 

(31) Commercial cut-off in per cent of stroke per cent 

(32) Initial pressure above atmosphere lb. per sq. in. 

(33) Back pressure at lowest point above or 

below atmosphere lb. per sq. in. 

(a) Mean back pressure above atmos- 
phere or zero lb. per sq. in. 

(34) Mean effective pressure lb. per sq. in. 

(a) Equivalent m.e.p. referred to 1st 

cylinder lb. per sq. in. 

(6) Equivalent m.e.p. referred to 2d 

cylinder lb. per sq. in. 

(c) Equivalent m.e.p. referred to 3d 

cylinder lb. per sq. in. 

(35) Aggregate m.e.p. referrearto each cylin- 

der lb. per sq. in. 

(36) Steam accounted for per i.hp-hr. at 

point on expansion line shortly after 

cut-off lb. 

(37) Steam accounted for per i.hp-hr. at 

point on expansion line just before 

release lb. 

(a) Pressure at selected point near 

cut-off t lb. per sq. in. 

(6) Pressure at selected point near 

release lb. per sq. in. 

(c) Pressure at point on compression 

curve shortly after exhaust closure lb. per sq. in. 

{d) Proportion of direct stroke com- 
pleted at selected point near cut- 
off 

(e) Proportion of direct stroke com- 
pleted at selected point near re- 
lease 

(/) Proportion of return stroke uncom- 
pleted at selected point on com- 
pression line 



1st 2d 3d 
Cyl. Cyl. Cyl 



In multiple expansion engines. 



t Pressures all referred to zero. 



APiPENDIX B 1015 



(g) Ratio of expansion 

(h) M.e.p. of hypothetical diagram 

(App. 27) lb. per sq. in. 

(i) Diagram factor (App. 27) 



Speed. 

(38) Revolutions per minute r.p.m. 

(39) Piston speed per minute ft. 

(a) Variation of speed between no load and full load per cent 

(6) Momentary fluctuation of speed on suddenly changing 

from full load to half-load per cent 

Power. 

(40) Indicated hp. developed, whole engine i.hp. 

(a) I.hp. developed by 1st cylinder i.hp. 

(b) I.hp. developed by 2d cylinder i.hp. 

(c) I.hp. developed by 3d cylinder i.hp. 

(41) Brake hp br. hp. 

(42) Friction of engine (Item 40 — Item 41) .hp. 

(a) Friction expressed in percentage of i.hp. (Item 42 -i- Item 

40 X 100) per cent 

(6) Indicated hp. with no load, at normal speed i.hp. 

Economy Results. 

(43) Dry steam consumed by engine per i.hp. per hr lb. 

(44) Dry steam consumed by engine per brake hp-hr lb. 

(45) Percentage of steam consumed by engine accounted for by 

indicator at point i)ear cut-off per cent 

(46) Percentage of steam consumed near release per cent 

(47) Heat units consumed by engine per i. hp-hr. (Item 30 -f- 

Item 40) B.t.u. 

(48) Heat units consumed by engine per br. hp-hr. (Item 30 -^ 

Item 31) B.t.u. 

Efficiency Results. 

(49) Thermal efficiency~of engine referred to i.hp. [(2546.5 -r- Item 

47) X 100] per cent 

(50) Thermal efficiency of engine referred to br. hp. [(2546.5 -j-Item 

48) X 100] per cent 

(51) Efficiency of Rankine cycle between temperatures of Items 20 and 21 . . 

(52) Rankine cycle ratio referred to i.hp. (Item 49 -j- Item 51) 

(53) Rankine cycle ratio referred to br. hp. (Item 50 -^ Item 51) 

Work Done per Heat Unit. 

(54) Net work per B.t.u. consumed by engine (1,980,000 -^ Item 48) ... . Ft-lb. 



1016 STEAM POWER PLANT ENGINEERING 

Sample Diagrams. 

(55) Sample diagrams from each cylinder 

(a) Steam pipe diagrams. 

Note: — For an engine driving an electric generator the form should be enlarged 
to include the electrical data, embracing the average voltage, number of amperes 
each phase, number of watts, number of watt hours, average power factor, etc.; 
and the economy results based on the electric output embracing the heat units 
and steam consumed per electric hp-hr. and per kw-hr,, together with the efficiency 
of the generator. (See table for Steam Turbine Code, Appendix C.) 

Likewise, in a marine engine having a shaft dynamometer, the form should in- 
clude the data obtained from this instrument, in which case the brake hp. becomes 
the shaft hp. 

PRINCIPAL DATA AND RESULTS OF RECIPROCATING ENGINE TEST. 

(1) Dimensions of cylinders 

(2) Date 

(3) Duration hr. 

(4) Pressure in steam pipe near throttle by gage lb. per sq. in. 

(5) Pressure in receivers lb. per sq. in. 

(6) Vacuum in condenser in. of mecury 

(7) Percentage of moisture in steam near throttle or number 

of degrees of superheating per cent or deg. 

(8) Net steam consumed per hour lb. 

(9) Mean effective pressure in each cylinder lb. per sq. in. 

(10) Revolutions per minute r.p.m. 

(11) Indicated horsepower developed i.hp. 

(12) Steam consumed per i.hp-hr lb. 

(13) Steam accounted for at cut-off each cylinder lb. 

(14) Heat consumed per i.hp-hr B.t.u. 



APPENDIX C 

DATA AND RESULTS OF STEAM TURBINE OR TURBO- 
GENERATOR TEST 

A.S.M.E. Code of 1915 

(1) Test of turbine located at 

To determine " 

Test conducted by 



Dimensions, etc. 

(2) Type of turbine (impulse, reaction, or combination) 

(a) Number of stages 

(6) Condensing or non-condensing 

(c) Diameter of rotors 

(d) Number and type of nozzles 

(e) Area of nozzles 

(J) Type of governor 

(3) Class of service (electric, pumping, compressor, etc.) 

(4) Auxiliaries (steam or electric driven) 

(a) Type and make of condensing equipment 

(6) Rated capacity of condensing equipment 

(c) Type of oil pumps (direct or independently driven) 

(d) Type of exciter (direct or independently driven) 

(e) Type of ventilating fan, if separately driven 

(5) Rated capacity of turbine 

(a) Name of builders 

(6) Capacity of generator or other apparatus consuming power of turbine. . . 

Date and Duration. 

(7) Date : 

(8) Duration hr. 

Average Pressures and Temperatures. 

(9) Pressure in steam pipe near throttle by gage lb. per sq. in. 

(10) Barometric pressure in. of mercury 

(a) Pressure at boiler by gage lb. per sq. in. 

(6) Pressure in steam chest by gage lb. per sq. in. 

(c) Pressure in various stages lb. per sq. in. 

(11) Pressure in exhaust pipe near turbine, by gage lb. per sq. in. 

1017 



1018 STEAM POWER PLANT ENGINEERING 

(12) Vacuum in condenser in. of mercury 

(a) Corresponding absolute pressure lb. per sq. in 

(6) Absolute pressure in exhaust chamber of turbine lb. per sq, in 

(13) Temperature of steam near throttle deg 

(a) Temperature of saturated steam at throttle pressure deg 

(6) Temperature of steam in various stages, if superheated deg 

(14) Temperature of steam in exhaust pipe near turbine deg 

(o) Temperature of circulating water entering condenser deg 

(6) Temperature of circulating water leaving condenser deg 

(c) Temperature of air in turbine room deg 

Quality of Steam. 

(15) Percentage of moisture in steam near throttle, or number 

of degrees of superheating per cent or deg. 

Total Quantities. 

(16) Total water fed to boilers lb. 

(17) Total condensate from surface condenser (corrected for condenser 

leakage and leakage of shaft and pump glands) lb. 

(18) Total dry steam consumed (Item 16 or 17 less moisture in steam). . . .lb. 

Hourly Quantities. 

(19) Total water fed to boilers or drawn from surface condenser per hour. .lb. 

(20) Total dry steam consumed for all purposes per hour (Item 18-4- 

Item 8) lb. 

(21) Steam consumed per hour for all purposes foreign to the turbine 

(including drips and leakage of plant) lb. 

(22) Dry steam consumed by turbine per hour (Item 20 — Item 21) lb. 

(a) Circulating water supplied to condenser per hour lb. 

Hourly Heat Data. 

(23) Heat units consumed by turbine per hour [Item 22 X (total heat 

of steam per pound at pressure of Item 9 less heat in 1 lb. of 

water at temperature of Item 14)] B.t.u. 

(a) Heat converted into work per hour B.t.u. 

(6) Heat rejected to condenser per hour (Item 22a X [Item 14& — 

Item 14a]) (approximate) B.t.u. 

(c) Heat rejected in the form of steam withdrawn from the turbine ... B.t.u. 

(d) Heat lost by radiation from turbine, and unaccounted for B.t.u. 

Electrical Data. 

(24) Average volts, each phase volts 

(25) Average amperes, each phase amperes 

(26) Average kilowatts, first meter kw. 

(27) Average kilowatts, second meter kw. 

(28) Total kilowatts output kw. 



APPENDIX C 1019 

(29) Power factor 

(30) Kilowatts used for excitation and for separately driven ventilating 

fan kw. 

(31) Net kilowatt output kw. 

Speed. 

(32) Revolutions per minute r.p.m. 

(33) Variation of speed between no load and full load r.p.m. 

(34) Momentary fluctuation of speed on suddenly changing from full 

load to half -load r.p.m. 

Power. 

(35) Brake horsepower, if determined br. hp. 

(36) Electrical horsepower e-hp. 

Economy Results, 

(37) Dry steam consumed by turbine per br. hp-hr lb. 

(38) Dry steam consumed per net kw-hr lb. 

(39) Heat units consumed by turbine per br. hp-hr. (Item 23 -i- 

Item 35) B.t.u. 

(40) Heat units consumed per net kw-hr B.t.u. 

Efficiency Results. 

(41) Thermal efficiency of turbine (2546.5 -^ Item 39) X 100 per cent 

(42) Efficiency of Rankine cycle between temperatures of Items 13 

and 14 "..... per cent 

(43) Rankine cycle ratio (Item 41 -r- Item 42) 

Work Done per Heat Unit. 

(44) Net work per B.t.u. consumed by turbine (1,980,000 -i- Item 39) ft.lb. 

PRINCIPAL DATA AND RESULTS OF TURBINE TEST. 

(1) Dimensions 

(2) Date 

(3) Duration hr. 

(4) Pressure in steam pipe near throttle by gage lb. per sq. in. 

(5) Vacuum in condenser in. of mecury 

(6) Percentage of moisture in steam near throttle or number 

of degrees of superheating per cent or deg. 

(7) Net steam consumed per hour lb. 

(8) Revolutions per minute r.p.m. 

(9) Brake horsepower developed br. hp. 

(10) Kw. output kw. 

(11) Steam consumed per brake hp-hr lb. 

(12) Heat consumed per brake hp-hr B.t.u. 

(13) Steam consumed per kw-hr lb. 

(14) Heat consumed per kw-hr B.t.u. 



APPENDIX D 

DATA AND RESULTS OF STEAM PUMPING MACfflNERY TEST 

A.S.M.E. Code of 1915 

(1) Test of pump located at 

To determine 

Test conducted by 



Dimensions, etc. 

(2) Type of machinery 

(3) Rated capacity in gallons per 24 hr gal. 

(4) Size of engine or turbine 

(5) Size of pump 

(6) Auxiliaries (steam or electric driven) 

(a) Type and make of condenser equipment '. . . . 

(6) Rated capacity of condenser equipment 

(c) Type of oil pump, jacket pump, and reheater pump (direct 

or independently driven) 

Date and Duration. 

(7) Date 

(8) Duration hr. 

Average Pressures and Temperatures 

(9) Pressure in steam pipe near throttle by gage lb. per sq. in. 

(10) Barometric pressure in. of mercury 

(a) Steam chest pressure lb. per sq. in. 

(6) Pressure in receivers and reheaters by gage lb. per sq. in. 

(c) Pressiu-e in turbine stages by gage lb. per sq. in. 

(11) Pressure in exhaust pipe near engine or turbine by gage lb. per sq. in. 

(12) Vacuum in condenser in. of mercury 

(a) Corresponding absolute pressure lb. per sq. in. 

(6) Absolute pressure in exhaust chamber lb. per sq. in. 

(13) Temperature of steam, if superheated, at throttle deg. 

(o) Normal temperature of saturated steam at throttle pressure deg. 

(6) Temperature of steam leaving receivers, if superheated deg. 

(14) Temperature of steam in exhaust pipe near engine or turbine deg. 

(a) Temperature of circulating water entering condenser deg. 

(6) Temperature of circulating water leaving condenser deg. 

1020 



APPENDIX D 1021 

(15) Pressure in force main by gage lb. per sq. in. 

(16) Vacuum or pressure in suction main by gage. 

in. of mercury or lb. per sq. in. 

(a) Correction for difference in elevation of the two gages . . lb. per sq. in. 

(17) Total head expressed in lb. pressure per sq. in lb. per sq. in. 

(a) Total head expressed in ft ft. 

Quality of Steam. 

(18) Percentage of moisture in steam near throttle, or number of degrees of 

superheating per cent or deg. 

Total Quantities. 

(19) Total water fed to boilers lb. 

(20) Total condensed steam from surface condenser (corrected for con- 

denser leakage) lb. 

(21) Total dry steam consumed (Item 19 or 20 less moisture in steam). . . .lb. 

(22) Total water discharged, by measurement gal. 

(a) Total water discharged, by plunger displacement, uncorrected. . . .gal. 

,, , ^ . c y fltem 22a - Item 22 ^ ^ „_-| 
(6) Percentage of slip [_ ^^^^^^-^2^ X lOOj 

(c) Leakage of pump gal. 

(d) Total water discharged, by calculation from plunger displacement, 

corrected for leakage gal. 

(e) Total weight of water discharged, as measured lb. 

(/) Total weight of water discharged, by calculation from plunger 

displacement, corrected for leakage lb. 

Hourly Quantities. 

(23) Total water fed to boilers or drawn from surface condenser per hour. . .lb. 

(24) Total dry steam consumed for all purposes per hour (Item 21 

^ Item 8) lb. 

(25) Steam consumed per hour for all purposes foreign to main engine lb. 

(26) Dry steam consumed by engine or turbine per hour (Item 24 — 

Item 25) lb. 

(a) Circulating water supplied to condenser per hour lb. 

(27) Weight of water discharged per hour, by measurement lb. 

(a) Weight of water discharged per hour, calculated from plunger 

displacement, corrected lb. 

Hourly Heat Data. 

(28) Heat units consumed by engine or turbine per hour [Item 26 

X (total heat of one lb. of steam at pressure of Item 9, 1/ess 

heat in one lb. of water at temperature of Item 14)] B.t.u. 

Indicator Diagrams. 

(29) Mean effective pressure, each steam cylinder y. . lb. per sq. in. 

(o) Mean effective pressure, each water cylinder, if any. . . .lb. per sq. in. 



1022 STEAM POWER PLANT ENGINEERING 

Speed and Stroke. 

(30) Revolutions per minute r.p.m. 

(a) Number of single strokes per minute strokes 

(6) Average length of stroke ft. 

Power. 

(31) Indicated horsepower developed i.hp. 

(a) Brake horsepower consumed by pump 

(32) Water horsepower hp. 

(33) Friction horsepower (Item 31 — Item 32) hp. 

(34) Percentage of i.hp. lost in friction per cent 

Capacity. 

(35) Water discharged in 24 hr., as measured gal. 

(o) Water discharged in 24 hr., calculated from plunger displacement, 

corrected gal. 

(b) Water discharged per minute, as measured gal. 

(c) Water discharged per minute, calculated from plunger displace- 

ment, corrected gal. 

Economy Results. , 

(36) Heat units consumed per i.hp-hr B.t.u. 

(37) Heat units consumed per water hp-hr B.t.u. 

(a) Dry steam consumed per i.hp-hr lb. 

(6) Dry steam consumed per water hp-hr lb. 

Efficiency Results. 

(38) Thermal efficiency referred to i.hp. [(2546.5 H- Item 36) X 100] . .percent 

(a) Thermal efficiency referred to water hp. [(2546.5 -^ Item 37) 

X 100] per cent 

(6) Mechanical efficiency y-^ ^ X 100 per cent 

(c) Pump efficiency r: oT~ X 100 per cent 

Duty. 

(39) Duty per 1,000,000 heat units ft-lb. 

Work Done per Heat Unit. 

(40) Work per B.t.u. (1,980,000 ^ Item 37) ft-lb. 

Sample Diagrams. 

(41) Sample indicator diagrams from each steam and pump cylinder 

Note: — The items relating to indicator diagrams and indicated horsepower are 
to be used only in the case of reciprocating machines. 



APPENDIX E 

DATA AND RESULTS OF STEAM POWER PLANT TEST 
A.S.M.E. Code of 1915 

(1) Test of plant located at 

To determine 

Test conducted by 



Date, Duration, etc. 

(2) Number and kind of boilers (superheaters, if any), engines, turbines, etc. 

(3) Rated capacity of boilers in lb. of steam per hour from and at 212 deg. lb. 

(a) Kind of furnace 

(6) Grate surface sq. ft. 

(c) Percentage of area of openings to area of grate per cent 

(d) Water heating surface sq. ft. 

(e) Superheating surface sq. ft. 

(4) Rated power of engines or turbines 

(a) Dimensions of cylinders of engine 

(6) Dimensions of turbine 

(c) Type of engines or turbines and class of service 

(d) Name of builders 

(5) Type of auxiliaries * 

(a) Dimensions of auxiliaries* 

(6) Type and capacity of condenser 

(7) Capacity of generators, pumps, or other apparatus consuming power of 

engine or turbine 

Date, Duration, etc. 

(8) Date 

(9) Duration. Length of time engine or turbine was in motion with 

throttle open hr. 

(a) Length of time engine or turbine was running at normal speed. . . .hr. 
(6) Elapsed time from start to finish hr. 

(10) Kind and size of coal 

* For full particulars see text of Report. 
1023 



1024 STEAM POWER PLANT ENGINEERING 

Average Pressures, Temperatures, etc. 

(11) Boiler pressure by gage lb. per sq. in. 

(a) Steam pipe pressure near throttle by gage lb. per sq. in. 

(6) Barometric pressure in. of mercury 

(c) Steam chest pressure by gage lb. per sq. in. 

(d) Pressure in receivers and reheaters by gage lb. per sq. in. 

(e) Pressure in turbine stages by gage lb. per sq. in. 

(/) Pressure in exhaust pipe near engine or turbine lb. per sq. in. 

(12) Vacuum in condenser in. of mercury 

(a) Corresponding absolute pressure lb. per sq. in. 

(6) Absolute pressure in exhaust chamber lb. per sq. in. 

(13) Temperature of steam, if superheated (taken at boiler or super- 

heater) deg. 

(a) Temperature of steam, if superheated (taken at throttle) deg. 

(b) Normal temperature of saturated steam at boiler pressure deg. 

(c) Normal temperature of saturated steam at throttle pressure deg. 

(d) Temperature of steam leaving receivers, if superheated deg. 

(e) Temperature of steam in exhaust pipe near engine or turbine .... deg. 

(/) Temperature of condensed water in hot-well or feed tank deg. 

(g) Temperature of circulating water entering condenser deg. 

(h) Temperature of circulating water leaving condenser deg. 

(i) Temperature of air in boiler room deg. 

(J) Temperature of air in engine or turbine room deg. 

(14) Temperature of feed water entering boilers (average) deg. 

(o) Temperature of each feed supply (if more than one) deg. 

(6) Temperature of feed water entering economizer, if any .deg. 

(c) Increase in temperature of water due to economizer deg. 

(15) Temperature of escaping gases leaving boiler deg. 

(a) Temperature of escaping gases leaving economizer deg. 

(6) Decrease in temperature of gases due to economizer deg. 

(c) Temperature of furnace deg. 

(16) Force of draft in main boiler flue in. of water 

(o) Force of draft at base of chimney in. of water 

(6) Force of draft at each end of economizer in. of water 

(c) Force of draft at individual boiler dampers in. of water 

(d) Force of draft in individual furnaces in. of water 

(e) Force of draft or blast in individual ash pits* in. of water 

(17) State of weather 

(a) Temperature of external air deg. 

Quality of Steam. 

(18) Percentage of moisture in steam, or number of degrees of 

superheating per cent or deg. 

(a) Factor of correction for quality of steam 

* If artificial draft or blast is employed, the force of draft or blast at the fan should 
also be given. 



APPENDIX £ 1025 

Total Quantities of Coal and Water. 

(19) Total weight of coal as fired lb. 

(a) Percentage of moisture in coal per cent 

(6) Total weight of dry coal . lb. 

(c) Total ash, clinkers, and refuse (dry) lb. 

(e) Percentage of ash and refuse in dry coal per cent 

(J) Total combustible burned (Item 196 - 19c) lb. 

(20) Total weight of water fed to boiler from all sources* lb. 

(o) Total water evaporated corrected for quality of steam (Item 20 

X Item 18a) lb. 

(6) Factor of evaporation based on average temperature of water en- 
tering boiler 

(c) Total equivalent evaporation from and at 212 degrees (Item 20a 

X Item 206) lb. 

(21) Coal, as fired, per hour (Item 19 -^ Item 9) lb. 

(a) Dry coal per hour (Item 196 -=- Item 9) lb. 

(6) Dry coal per sq. ft. of grate surface lb. 

(22) Water evaporated per hour (Item 20 -r- Item 9) lb. 

(a) Equivalent evaporation per hour from and at 212 deg lb. 

(6) Equivalent evaporation per sq. ft. of water heating surface lb. 

(23) Dry steam generated per hour (sum of sub-items a to g) (Item 20 

less moisture in steam -^ Item 9) lb. 

(a) Moisture formed per hour between boiler and engine lb. 

(6) Dry steam consumed per hour by engine cylinders or turbine lb. 

(c) Dry steam consumed per hour by reheaters and jackets, if any. . . .lb. 

(d) Dry steam consumed per hour by air and circulating pump of con- 

denser lb. 

(e) Dry steam consumed per hour by boiler-feed pump lb. 

(J) Dry steam consumed per hour by other steam-driven auxiliaries. . .lb. 
(g) Dry steam consumed per hour to supply leakage of boilers and 

piping between boilers and engine (including steam supplied for 
foreign purposes, if any) lb. 

(h) Live steam suppHed for heating, or miscellaneous purposes lb. 

(i) Injection or circulating water supplied condenser per hour lb. 



Calorific Value of Coal. 

(24) Calorific value of 1 lb. of coal as fired, by calorimeter test B.t.u. 

(a) Calorific value of 1 lb. of dry coal B.t.u. 

(6) Calorific value of 1 lb. of combustible B.t.u. 

* If there are a number of supplies of feed water, the weight and temperature of 
each supply is to be given, and total weight and average temperature ascertained. 



1026 STEAM POWER PLANT ENGINEERING 

Hourly Heat Data. 

(25) Heat units in coal as fired generated per hour (Item 21 X Item 24) .B.t.u. 

(26) Heat units consumed by engine and auxiliaries per hour (Item 

22 X total heat of 1 lb. of steam at pressure of Item 11 less heat 
in 1 lb. of water at temperature of feed water supplied to boiler, 
or economizer, if any) B.t.u. 

(a) Heat converted into work per hour B.t.u. 

(b) Heat rejected to condenser per hour B.t.u. 

(c) Heat rejected in steam withdrawn from receivers or turbine- 

stages not used by feed water ' B.t.u. 

(d) Heat lost by radiation from engine and auxiUaries, including 

piping between boilers and condenser B.t.u. 

(e) Heat lost in operation of boiler, including economizer (if any) 

(Item 25 - Item 26) B.t.u. 

Indicator Diagrams. 

(27) Mean effective pressure, each cylinder lb. 

(a) Commercial cut-off (in per cent of stroke) each cylinder per cent 

(6) Initial pressure, above atmosphere, each cylinder lb. per sq. in. 

(c) Back pressure at lowest point above or below atmosphere, 

each cylinder lb. per sq. in. 

(d) Steam accounted for per i.hp. per hour at point near cut-off, 

each cylinder lb. 

(e) Steam accounted for per i.hp. per hour at point near release lb. 

Electrical Data. 

(28) Average kilowatt output, gross kw. 

(a) Volts each phase volts 

(6) Amperes each phase amperes 

(c) Kilovolt amperes kv-a. 

(d) Power factor 

(29) Current used by exciter kw. 

(30) Net kilowatt output (Item 28 - Item 29) kw. 

(31) Revolutions per minute r.p.m. 

(a) Variation of speed between no load and full load r.p.m. 

Power. 

(32) Indicated horsepower i.hp. 

(33) Brake horsepower br. hp. 

Capacity. 

(34) Water evaporated per hour from and at 212 degrees (same as Item 

22a) lb. 

(a) Percentage of rated boiler capacity developed (Item 34 -i- Item 

3 X 100) per cent 



APPENDIX E 1027 

(35) Percentage of rated engine or turbine capacity developed (Item 

32 -T- Item 4 X 100) per cent 

Economy Results. 

(36) Coal as fired per i.hp. of engine per hour lb. 

(37) Coal as fired per brake hp. of engine or turbine per hour lb. 

(a) Dry coal per i.hp. per hr lb. 

(6) Dry coal per brake hp-hr lb. 

(c) Dry coal per kw-hr lb. 

(38) Heat units in coal consumed per i.hp. of engine per hour B.t.u. 

(39) Heat units in coal consumed per brake hp. of engine or turbine 

per hour (Item 37 X Item 24) B.t.u. 

(a) Heat units consumed by engine (including auxiliaries) per 

i.hp-hr . . B.t.u. 

(6) Heat units consumed by engine or turbine (including auxiliar- 
ies) per brake hp-hr. (Item 26 -r- Item 33) B.t.u. 

(c) Heat units consumed by engine per kw-hr B.t.u. 

(40) Heat units in coal consumed per kw-hr B.t.u. 

(41) Water evaporated per lb. of coal as fired lb. 

(a) Water evaporated per lb. of dry coal lb. 

(6) Equivalent evaporation from and at 212 deg. per lb. of dry coal. . .lb. 
(c) Equivalent evaporation from and at 212 deg. per lb. of combustible, .lb. 

(42) Dry steam consumed by engine alone per i.hp-hr lb. 

(a) Dry steam consumed by auxiliaries per i.hp-hr lb. 

(b) Dry steam consumed by combmed engine and auxiUaries per i.hp-hr . lb. 

(43) Dry steam consumed by engine or turbine alone per brake hp-hr lb. 

(a) Dry steam consumed by auxiliaries per brake hp-hr lb. 

(6) Dry steam consumed by combined engine or turbine and auxiliaries 
per brake hp-hr lb. 

Efficiency Results. 

(44) Thermal efficiency of plant referred to i.hp. [(2546.5 -^ Item 38) 

X 100] 

(45) Thermal efficiency of plant referred to brake hp. [(2546.5 -J- Item 

39) X 100] 

(a) Efficiency of boilers (Item 416 X 970.4 X 100 -r- Item 24a) 

(6) Efficiency of engine referred to i.hp. [(2546.5 ^ Item 39a) X 100]. . . . 

(c) Efficiency of engine or turbine referred to brake hp. [(2546.5 -^ 396) 

X 100] 

Fuel Cost of Power. 

(46) Cost of coal per ton of lb dollars 

(47) Cost of coal per i.hp-hr cents 

(48) Cost of coal per brake hp-hr cents 



1028 



STEAM POWER PLANT ENGINEERING 

HEAT BALANCE OF STEAM POWER PLANT. 



(49) Heat units in coal (same as Item 24) 

(50) Boiler losses 

(a) Loss due to evaporation of moisture in coal 

(6) Loss due to heat carried away by steam formed by 
the burning of hydrogen , 

(c) Loss due to heat carried away in the dry flue gases 

(d) Loss due to carbon monoxide 

(e) Loss due to combustible in ash and refuse 

(/) Loss due to heating moisture in air 

Ig) Loss due to unconsumed hydrogen and hydrocar- 
bons, to radiation, and unaccounted for 

(h) Heat supplied steam-driven appliances for operat- 
ing boilers less that recovered by heating feed 
water .' 

(i) Total boiler losses 

(51) Engine consumption 

(a) Radiation from steam pipe 

(6) Radiation from engine or turbine 

(c) Heat rejected to condenser 

(d) Heat withdrawn from engine receivers or turbine 

stages or other use than heating feed water 

(e) Heat lost by leakage of steam piping 

(f) Heat converted into work 

(52) Heat in steam supplied for purposes foreign to engine 

or turbine 

Totals (same as Item 49) 



Per Lb. 
Coal as 
Fired. 



Per Cent. 



Sample Diagrams. 

(53) Sample indicator diagrams from each cylinder of en- 
gine. Also sample steam pipe diagrams 

The boiler output (Item 49— Item 50i) may be divided into 

(a) Heat units absorbed by water in boiler 

(b) Heat units absorbed by water in economizer 

The quantity representing the sum of Items 516, c, and / 

may be divided according to the steam distribution into. . . 

(c) Heat consumed by engine cylinders or turbine alone 

(including reheaters or jackets, if any), i.e., total 
heat supplied to engine or turbine alone less heat re- 
covered therefrom by heating feed water 

(d) Heat consumed by steam-driven auxiliaries, i.e., total 

heat supplied to auxiliaries less heat recovered there- 
from by heating feed water 

The same quantity may be divided according to the distri- 
bution of work done by engine or turbine into 

(e) Heat consumed in supplying power lost in friction of 

engine or turbine 

if) Heat consumed in supplying frictional, electrical, or 
other losses of power delivered by engine or turbine 
shaft 

ig) Heat consumed in supplying useful power delivered by 
engine or turbine, whether mechanical, electrical, or 
otherwise 



APPENDIX E 1029 

Note: — In the case of pumping and air machinery plants add lines under 
the various items as follows: 

For Item (13) 

(k) Pressure in delivery main by gage lb. per sq. in. 

(0 Vacuum or pressure in suction main 

by gage lb. per sq. in. or in. of mercury 

(m) Correction for difference in elevation of the two gages. . .lb. per sq. in. 
(n) Total head expressed in lb. pressure per sq. in lb. per sq. in. 

(0) Total head expressed in ft ft. 

For Item (20) 

(d) Temperature of delivery deg. 

(e) Total weight of water discharged, by measurement lb. 

(/) Total weight of water discharged, by calculation from plunger 

displacement, corrected lb. 

{g) Total volume of air delivered, by measurement cu. f t. 

(A) Total volume of air delivered, reduced to atmospheric pressure and 

temperature cu. ft. 

For Item (23) 

(j) Weight of water discharged per hour, by measurement lb. 

(/c) Weight of water discharged per hour, by plunger displace- 
ment, corrected lb. 

(1) Volume of water or air delivered per hour, by measurement cu. ft. 

(w) Volume of air delivered per hour, reduced to atmospheric pres- 
sure and temperature cu. ft. 

For Item (31) 

(6) Length of pump stroke ft. 

For Item (33) 

(a) Water (or air) hp hp. 

For Item. (35) 

(a) Gal. of water discharged in 24 hr. as measured gal. 

(6) Volume of air delivered per minute, reduced to atmospheric 

pressure and temperature cu. ft. 

For Item (36) 

(o) Dry coal per water (or air) hp-hr lb. 

For Item (39) 

(a) Duty per 1,000,000 B.t.u 

For Item (45) 

(a) Thermal efficiency of plant referred to water (or air) hp 

For Item (48) 

(a) Cost of coal per water (or air) hp dollars. 



1030 STEAM POWER PLANT ENGINEERING 



PRINCIPAL DATA AND RESULTS OF STEAM POWER PLANT TEST. 

(1) Dimensions of boilers 

(2) Dimensions of engine or turbine 

(3) Date 

(4) Duration hr. 

(5) Boiler pressure lb. per sq. in. 

(6) Throttle pressure lb. per sq. in. 

(7) Pressure in receiver or stages lb. per sq. in. 

(8) Vacuum in condenser in. of mercury 

(9) Percentage of moisture in steam near throttle or number 

of degrees of superheating per cent or deg. 

(10) Temperature of feed water entering boilers deg. 

(11) Temperature of escaping gases deg. 

(12) Force of draft in. of water 

(13) Coal, as fired, per hour lb. 

(14) Percentage of moisture in coal per cent 

(15) Percentage of ash in coal per cent 

(16) Water evaporated per hour lb. 

(17) Equivalent evaporation per hour from and at 212 deg lb. 

(a) Equivalent evaporation per hour from and at 212 deg. 

per sq. ft. water heating surface lb. 

(18) Steam consumed per hour by engine lb. 

(19) Steam consumed per hour by engine or turbine and auxiliaries lb. 

(20) Mean effective pressure in each cylinder of engine lb. per sq. in. 

(21) Revolutions per minute - r.p.m. 

(22) Indicated horsepower i.hp. 

(23) Brake horsepower * brake hp. 

(24) Coal as fired per i.hp-hr lb. 

(25) Coal as fired per brake hp-hr. * lb. 

(26) Steam per i.hp-hr lb. 

(27) Steam per brake hp-hr.* lb. 

(28) Heat consumed per i.hp-hr. . •. B.t.u. 

(29) Heat consumed per brake hp-hr.* B.t.u. 

* For pumping engine (water or air) use Water or Air hp. in place of Brake hp. 



APPENDIX F 



MISCELLANEOUS CONVERSION FACTORS 



1 Pound per Square Inch = 

2.0355 inches of mercury at 32° F. 
2.0416 inches of mercury at 62° F. 
2.309 feet of water at 62° F. 
0.07031 kilogram per square centi- 
meter 
0.06804 atmosphere 
51.7 milUmeters of mercury at 

32° F. 

1 Foot of Water at 62° F. = 
0.433 pound per square inch 
62.355 pounds per square foot 
0.883 inch of mercury at 62° F. 
821.2 feet of air at 62° F. and ba- 

rometer 29.92 

1 Inch of Water 62° F. = 
0.0361 pound per square inch 
5.196 pounds per square foot 
0.5776 ounce per square inch 
0.0736 inch of mercury at 62° F. 
68.44 feet of air at 62° F. and ba- 
rometer 29.92 

1 Foot of Air at 32° F. and Barom- 
eter 29.92 = 
0.0761 pound per square foot 
0.0146 inch of water at 62° F. 

1 Inch of Mercury at 62° F. = 
0.4912 pound per square inch 
1.132 feet of water at 62° F. 
13.58 inches of water at 62° F. 



1 Atmosphere = 

760.0 milUmeters of mercury at 

32° F. 
14.7 pounds per square inch 
29.921 inches of mercury at 32° F. 
2116.0 pounds per square foot. 
1.033 kilograms per square centi- 
meter 

1 Millimeter = 0.03937 inch 

1 Centimeter = 0.3937 inch 

1 Meter = 39.37 inches 

1 Meter = 3.2808 feet 

1 Square Meter = 10.764 square feet 

1 Liter = 
61.023 cubic inches 
0.264 U. S. gallons 

1 Gram = 

1 cubic centimeter of distilled 
water 
15.43 grains troy 
0.0353 ounce 

1 Kilogram = 
2,20462 pounds avoirdupois 



1031 



APPENDIX G 

EQUIVALENT VALUES OF ELECTRICAL AND MECHANICAL UNITS 



1 Myriawatt = 

10 kilowatts 
10,000 watts 

13.41 horsepower 
13.597 cheval-vapeur 
13.597 pferde-kraft 
26,552,000 foot pounds per hour 
8,605,000 gram calories per hour 
3,670,000 kilogram meters per hour 
34,150 B.t.u. per hour 
1.02 boiler horsepower 

1 Horsepower = 
745.7 watts 
0.7457 kilowatt 
0.07457 myriawatt 
1.0139 cheval-vapeur 
1.0139 pferde-kraft 
33,000 foot pounds per minute 
641,700 gram calories per hour 
273,743 kilogram meters per hour 
2,547 B.t.u. per hour 

1 Joule = 

1 watt second 
0.10197 kilogram meter 
0.73756 foot pound 
0.239 gram calorie 
0.0009486 B.t.u. 

1 B.T.U. = 

1054 watt seconds 
777.5 foot pounds 
107.5 kilogram meters 
0.0003927 horsepower hour 



1 Kilowatt = 

0.1 myriawatt 
1000 watts 
1.341 horsepower 
1.3597 cheval-vapeur 
1.3597 pferde-kraft 
2,655,200 foot pounds per hour 
860,500 gram calories per hour 
367,000 kilogram meters per hour 
3,415 B.t.u. per hour 
0.102 boiler horsepower 

1 Cheval-Vapeur or Pferde-Kraft 
75 kilogram meters per second 
0.07354 myriawatt 
0.7357 kilowatt 
0.9863 horsepower 
32,550 foot pounds per minute 
632,900 gram calories per hour 
2,512 B.t.u. per hour 

1 Foot Pound = 

1.3558 joules 
0.13826 kilogram meter 
0.001286 B.t.u. 
0.03241 gram calorie 
0.000000505 horsepower hour 

1 Kilogram-Meter = 
7.233 foot pounds 
9.806 joules 
2.344 gram calories 
0.0093 B.t.u. 



1032 



INDEX 



Accessories, boiler, 177-185. 
Accumulator, regenerative, 483. 
Acetylene, fuel properties of, 43. 
Acids in lubricating oils, 782. 
Acme bucket trap, 692. 
Acton relief valve, 773. 
Adiabatic change of state, 974. 
Admiralty metal condenser tubes, 631. 
Amortization, 868. 
Air, Carrier's chart, 996. 

chambers, 629. 

composition of, 47. 

cooled condensers, 541. 

dehumidifying, 1002. 

density of (table), 280. 

dew point, 998. 

drying loss, 24. 

heat loss due to excess, 61. 

leakage, condenser, 504-655. 

lift, 676. 

manometer, 826. 

moisture in, 994-999. 

nitrogen in, 47. 

properties of, 991-1006. 

dry, 991. 

partially saturated, 998. 

saturated, 994. 

pumps, 651-662 (see Pumps, vacuum). 

requirements for combustion, 47, 92. 

for fuel oil, 92. 

solid fuels, 55. 

space in grate bars, 176. 
in boiler settings, 127. 

theoretical combustion requirements, 
47. 

weight per lb. various fuels, 55. - 
Alarm, high and low water, 180. 
Alberger condenser, 619. 
Alkalinity, 577. 
Allis-Chalmers turbine, 473. 
Analyses, coal, 23. 

feed water, 567. 



Analyses, flue gas, 63. 

fuel oil, 90. 

lubricating oil, 781. 

proximate, 23. 

scale, 567. 

ultimate, 25. 

visible smoke, 188. 
Anchor bolts, steel stacks, 306. 
Anchors, pipe, 728. 
Anderson feed- water purifier, 571. 
Animal fats, 777. 
Annuities, 867. 
Apron conveyors, 254. 
Aqueous vapor in condensers, 505. 
Arches, deflecting, 196. 
Areas of chimneys, 291. 
Argand blower, 328. 
Arithmetic mean temperature, 537. 
Armour Institute, chimney at, 310. 
Ash, combustible in, 66. 

composition of, 24. 

cost of handling, 273. 

fusibility of, 77. 

influence on fuel value of coal, 76. 
Ash bins, 250. 

handling systems, 248-274. 

hoppers, 10. 

pit, 1. 
A. S. M. E. testing codes, 1007-1020. 

boiler, 1007. 

power plants, 1023. 

pumping engines, 1020. 

reciprocating engines, 1012. 

turbines and turbo-generators, 1017 
Attendance, 871. 
Atmospheric condenser, 543. 

feed-water heater, 586. 

heaters, 15, 586. 

relief valve, 8, 773. 

surface lubrication, 789. 
Augmenter, Parsons vacuum, 661. 
Austin separator, 683. 



1033 



1034 



INDEX 



Automatic cut-off, 413. 

damper control, 179. 

injectors, 647. 

non-return valve, 765. 

relief valves, 8, 772. 

temperature control, 752. 
Auxiliaries, 8. 
Available draft, 284. 

hydrogen, 27. 
Avogadro's law, 52. 

Babcock & Wilcox boilers, 129. 

chain grates, 203. 

superheaters, 229. 
Back connection, boiler, 128. 
Back-pressure, decreasing, 383. 

valves, 5, 773. 
Badenhausen boiler, 139-610. 
Baffles, 8, 196. 
Bagasse as fuel, 38. 
Bagasse furnace, 220. 
Bailey steam meter, 824. 
Balanced draft, 327. 
Banking, coal burned in, 72. 

fires, hand-fired furnaces, 192. 
Baragwanath condensers, 515-523. 

feed water heater, 593. 
Barnard- Wheeler cooling tower, 561, 
Barometric condenser, 517. 

readings, corrections for, 501. 
Baum, oil separator, 685. 
Baum^ gravity, 89. 
Bearings, lubrication of, 789. 
Bellis-Morcom engine test, 388. 
Belt conveyors, 263. 
Bends, pipe, 726. 
Bernouilli's theorem, 757. 
Bigelow-Hornsby boiler, 135. 
Billow fuel-oil burner, 135. 
Binary-vapor engine, 398. 
Bins, coal-storage, 250. 
Bituminous coals, 35. 
Blades, turbine, 430-470. 
Bladeless turbine, 498. 
Blake jet condenser, 510. 

powdered-coal furnace, 89. 
Blast furnace gas, 110. 
Bleeder turbine, 486. 
Blonck efficiency meter, 826. 
Bloomsburg jet, 328. 
Blowers, centrifugal, 330-398 {see Fans). 



Blowers, soot, 181. 
Blowing off, loss due to, 73. 
Blow-off cocks, 4, 767. 

piping, 769. 

tank, 177. 

valves, 767. 
Blow-offs, 177. 
Boiler, boilers, 115-185. 

A. S. M. E. Code, 1007. 

Babcock & Wilcox, standard, 129. 

Babcock & Wilcox, cross-drum, 13L 

Badenhausen, 139, 610. 

Bigelow-Hornsby, 135. 

blast furnace, 336. 

blow-offs for, 177. 

builders rating, 150. 

capacity of, 163. 

classification, 115. 

combustion rates in, 151-3. 

Commonwealth Edison, 161. 

compounds for, 572. 

control boards, 843. 

corrosion in, 570. 

cost of, 172. 

cross-drum type, 131. 

curves of performance, 165-7. 

damper regulators, 178. 

Delray, performance of, 150. 

down-flow, 133. 

draft loss through, 285-7. 

dry pipe, 128. 

economical loading, 168. 

efficiency of, 154. 

efficiency-capacity relation, 165. 

factor of evaporation, 143. 

fire-box, 118. 

flue-gas temperatures in, 164. 

foaming in, 569. 

forced capacity of, 166. 

fuel oil for, 89-112. 

furnaces, 190-221. 

fusible plugs for, 181. 

grate surface for, 151. 

heat balance, 71. 

heat losses, 61-75. 

heating surface, 148. 

heat transmission, 144. 

height of chimney for, 298. 

Heine, 130. 

high-pressure, 139. 

high water alarm, 180. 



INDEX 



1035 



Boiler, horsepower of, 150. 

inherent losses, 73. 

Kewanee boiler, 118. 

losses, standby, 71. 

Manning, 117. 

Parker, 132. 

perfect boiler, efficiency of, 157. 

performance of, 158. 

rating of, 150. 

return-tubular, 122. 

Robb-Mumford, 121. 

safety valves, for, 768-772. 

Scotch-marine, 119. 

selection of type, 173. 

settings for, 122, 190-221. 

smoke prevention in, 190-221. 

soot removal from, 181. 

steam domes for, 128. 

Stu-ling, 136. 

superheated steam for, 223-249. 

temperature drop in, 172. 

tests of, 158-165. 

thickness of fire in, 169. 

tube cleaners, 183. 

unit of evaporation, 143. 

vertical-tubular, 116. 

Wickes, 134. 
Bolts, chimney foundation, 306. 
Bone, surface combustion, 112. 
Bourdon pressure gauge, 825. 
Boyle's and Charles' law, 503. 
Branch fuel-oil burner, 96. 
Breeching, 8, 321. 

Brick chimneys, 308-312 {see Chim- 
neys). 
Bridge wall, 8. 

Bridge wall, double-arch, 196. 
Bristol thermometers, 828. 
Brown coal, 37. 
Bucket conveyors, 257. 

traps, 682. 
Buckets, steam turbine, 430-470. 
Buckeye skimmer, 178. 
Buckeye locomobile, 408. 
Buffalo blowers, 337-347. 

General Electric Go's, plant, 924. 
Builders rating, 150. 
Building fires, hand-fired furnaces, 191. 
Bundy separator, 682. 

trap, 693. 
Bunkers, coal, 250. 



Burke's smokeless furnace, 191. 
Burners, fuel-oil, 94-7 (see Fuel oil). 

powdered coal, 81-7. 
Burning point of oils, 785. 
Bursting strength of pipes, 708. 

Cable car conveyors, 268. 
Caking coals, 35. 
Calorific value, coal, 44. 

fuel oil, 47. 
Capacity, boiler, 163. 

fans, 341-6. 

pipes, 710. 
Carbon, combustion data, 43. 

dioxide, combustion data, 43. 
index to combustion, 54. 
testing apparatus, 835. 

hydrogen ratio, 30. 

monoxide combustion data, 43. 

heat loss due to, 65. 

packing, 464. 
Carnot cycle, 971. 
Carpenter steam calorimeter, 841. 
Carrier's psychrometric chart, 996. 
Cast-iron pipe, 713. 
Cast-steel pipe, 707. 
Central condensing system, 554. 

oiling system, 798 {see Lubrication). 

station, elementary steam, 1-20. 
Buffalo General Electric, 924. 
Commonwealth Edison, 927. 
Essex, N. J., 914. 

statistics, 845-890. 
Centrifugal fans, 330-348 {see Fans). 

pumps, 664-9 (see Pumps, centrif- 
iigal). 
Chain grates, 203. 
Check valves, 767. 
Chemical analysis, feed water, 572. 

fuel oil, 89. 

fuels, 23-40. 

lubricating oils, 781. 

scale, 567. 
Chemical purification, feed water, 576. 
Chezy's formula, 288. 
Chicago smokeless settings, 194-9. 
Chimneys, 279-326. 

areas of, 291. 

breechings for, 321. 

brick, 308-312. 

at Armour Institute, 310. 



1036 



INDEX 



Chimneys, brick, core and linings for, 
312. 

cost of, 325. 

Custodis radial, 315. 

materials for, 312. 

radius of statical moments, 313. 

stability of, 313. 

thickness of walls, 308. 
classification of, 296. 
concrete (reinforced), 317-21. 

strain sheet, 321. 

Turneaure and Maurer curves, 320. 

Weber coniform, 318. 

Wiederholt, 317. 

wind stresses in, 320. 
cost of, 325. 

density of gases in, 280. 
draft in, 280. 
eccentricity in, 321. 
efficiencies of, 324. 
equations for design of, 295. 
evase, 334. 
foundations for, 322. 
general theory, 279. 
guyed, 299. 
height of, 297. 
horsepower rating of, 297. 
oil fuel, 296. 
stability of, 306, 313. 
steel, 299-308. 

foundations for, 306. 

foundation bolts for, 306. 

guyed, 299. 

plates for, 303. 

riveting for, 305. 

stability of, 306. 
wind pressure on, 302. 
Circulating pumps, 672. 
Classification of boilers, 115. 
chimneys, 298-308. 
coals, 30. 
condensers, 508. 
conveyors, 251. 
feed water heaters, 586. 
fuel oil burners, 94. 
fuels, 22. 

powdered coal furnaces, 82. 
pumps, 622. 
steam engines, 400. 

meters, 814. 

separators, 680. 



Classification of steam traps, 690. 
turbines, 424. 

stokers, 202. 

thermometers, 831. 

testing instruments, 805. 
Clayton's method, 987. 
Cleaner, tube, 91. 
Cleaning fires, 191. 
Clearance volume, 370. 
Closed heaters, 590-605. 
Coal, 23-56. 

analysis of, 23. 

analysis of various, 32-37. 

air drying loss, 24. 

air requirements for, 56. 

anthracite, 32. 

composition of, 32. 
sizes of, 33. 

ash in, 24. ' 

available hydrogen in, 27. 

bituminous, 35. 
composition of, 38. 
heating value of, 36. 
sizes of, 77. 

brown, 37. 

bunkers, 250. 

burned in banking, 72. 

caking and non-caking, 35. 

calorimeter, 842. 

calorimetric tests of, 32-38. 

classification of, 30. 

combined moisture in, 27. 

combustible in, 24. 

combustion of, 41. 

composition of, 23. 

conveyors, 248-274 (see Conveyors). 

cost of, for power, 872. 

draft requirements for, 284. 

dry, 25. 

Dulong's formula, 45. 

fixed carbon, 24. 

government specifications for pur- 
chasing, 909. 

handling machinery, 248-274. 

heat losses in burning, 60-75. 

heating values of, 44. 

hoppers, 274. 

hydrogen in, 26. 

meter, 804. 

moisture in, 23. 

nitrogen in, 49. 



INDEX 



1037 



Coal, powdered, 80-86 (see Powdered coal) . 

products of combustion, 49. 

proximate analysis, 23. 

selection and purchase, 72. 

semi-anthracite, 33. 

semi-bituminous, 34. 

spontaneous combustion, 250. 

storage, 249. 

sub-bituminous, 34. 

total moisture, 27. 

ultimate analysis, 25. 

valves, 276. 

volatile matter in, 24. 

washing, 79. 

weathering, 250. 
Cochrane heater, 587. 
Cocks, blow off, 767. 

try, 2, 180. 
Coefficient of expansion, pipes, 724. 

friction chimney gases, 288. 

steam in pipes, 744. 

water in pipes, 721. 

economizers, 612. 

feed-water heaters, 596. 

superheaters, 239. 
Coil heaters, 592. 
Coke oven gas, 110. 
Coking coals, 35. 
Cold process water purification, 577. 

test lubricating oils, 785. 
Color test, lubricating oils, 784. 
Column, water level, 180. 
Columbia steam trap, 693. 
Combined moisture in coal, 27. 
Combustible matter in coal, 24. 
Combustion, flameless, 112. 

products of, 49. 

rate of, 151. 

temperature of, 58. 

theory of, 41. 

surface, 112. 
Commercial efficiency, 1. 
Commonwealth-Edison, boiler test, 
161. 

coal and ash system, 261, 

condenser, 552. 

economizer test, 612. 

northwest Unit No. 3, 16, 927. 

pressure drop, steam main, 746. 
Compound engines, 405. 

steam turbines, 477, 490. 



Compounding, 392. 
Compounds, boiler, 572. 
Compressed air, lubrication, 794. 
Compression in engines, 371. 
Concrete chimneys, 317. 
Condensate, 12. 

pumps, 675. 
Condensation, cylinder, 366. 
Condensers, 499-565. 

acjueous vapor in, 505. 

air for cooling, 542. 

air leakage in, 537, 655. 

air pressure in, 504. 

air pumps for, 651. 

Alberger barometric, 519. 

arithmetic mean temperature in sur- 
face, 537. 

atmospheric, 543. 

Baragwanath surface, 515. 

barometric, 517. 

Blake jet, 510. 

classification of, 507. 

central, 554. 

centrifugal jet, 514. 

choice of, 556. 

C. H. Wheeler, low level, 513. 

circulating pumps for, 672-675. 

Commonwealth Edison, 50,000 sq. ft., 
552. 

cooling water for jet, 520. 

cooling water for surface, 527. 

cost of, 553. 

Dalton's law, 502. 

differential tube spacing, 526. 

dry air surface, 541. 

ejector, 516-8. 

exponential mean temperature, 537. 

evaporative surface, 545, 1003. 

heat transmission in surface, 529. 

hot well depression, 521-534. 

injection orifice, 511. 

injection water, 520. 

jet, 512-523. 

Alberger barometric, 519. 
Blake, 510. 
C. H. Wheeler, 513. 
cooling water for, 520. 
high vacuum, 512-4. 
low-level, 514. 
Tomlinson type, B, 520. 
Weiss counter-current, 518. 



1038 



INDEX 



Condensers, jet, Westinghouse-Leblanc, 
512. 

Wheeler, low-level, 514. 

Worthington low-level, 509. 
Koerting multi-jet, 517. 
location of, 547. 

logarithmic mean temperature dif- 
ference, 507. 
measurement of vacuum, 501. 
multi-jet, 517. 
Orrok, tests by, 534-7. 
parallel flow, 509. 
Pennel atmospheric, 543. 
power gained by use of, 506. 
pressure drop in surface, 526. 
proportions of surface, 540. 
radial flow surface, 526. 
saturated air, 543. 
Schutte ejector, 517. 
siphon, 515. 
surface, 523-547. 

air in, 504, 536, 655. 

air pumps for, 654. 

Baragwanath, 515. 

circulating pumps for, 672. 

circulating water for, 527. 

cleanliness of tubes, 531. 

coefficient of heat transfer in, 530. 

Commonwealth Edison, 50,000 sq. 
ft., 552. 

cooling water calculations, 527. 

cost of, 553. 

critical velocity of water in, 532. 

differential tube spacing, 526. 

dry air, 541. 

evaporative, 545. 

heat transmission in, 532-6. 

hot well depressions in, 534. 

Pennel saturated air, 534. 

proportions of, 540. 

pumping engines, 549. 

saturated air, 544. 

tests of evaporative, 546. 

tests of large surface, 540. 

Westinghouse radial flow, 512. 

Wheeler, admiralty, 535. 
surface, 526. 
saturated air, 544. 
tests of, dry air, 542. 

evaporative, 546. 

saturated air, 544. 



Condensers, tests of, surface, 540. 

Tomlinson barometric, 520. 

volume of condensing chamber, 511. 

water for, jet, 520. 
surface, 528. 

Weiss counter-current, 518. 

Westinghouse, Leblanc, 512. 
radial flow, 527. 

Wheeler, admiralty, 525. 
low-level jet, 514. 
rectangular jet, 551. 
surface, 526. 

Wheeler, C. H. high-vacuum, 550. 
low-level jet, 513. 

Worthington jet, 509. 
Condensing steam engines, 383. 

steam turbines, 496. 
Conoidal fans, 337-342. 
Control boards, boiler, 843. 
Conversion tables, 1032. 
Conveyors (coal and ash), 249-270. 

apron, 254. 

belt, 263. 

cable car, 268. 

classification of, 250. 

Commonwealth Edison, 261. 

continuous, 252. 

elevating tower, 266. 

flight, 253. 

hoist and trolley, 268. 

Hunt, 262. 

open top, 256. 

pan, 256. 

Peck, 258. 

pivoted bucket, 260. 

Robins, 264. 

scraper, 253. 

screw, 252 

telpherage, 268. 

vacuum, 270. 

V-bucket, 257. 
Cooling air, condensers, 541-560. 

humidifying, 998. 
Cooling ponds, 558. 

towers, 561, 1005. 

water, 520-560. 
Copper pipes, 707. 
Core and lining chimney, 312. 
Corliss engines, 359-418. 
Corrosion, 570. 
Cost of boilers and settings, 172. 



INDEX 



1039 



Cost of chimneys, 325. 

condensers, 555. 

engines, 416. 

handling coal and ashes, 273. 

mechanical draft, 349. 

pipe flanges, 719. 

power, 845-890 {see Power cost). 

power plants, 850-890. 

pulverizing coal, 87. 

stokers, 222. 

turbines, 486. 
Countercurrent flow, 612. 
Covering, pipe, 721. 
Critical temperature of steam, 944. 
Cross compound engines, 405. 
Cross-over main, 732. 
Curtis turbines, 453-467 (see Turbines, 

steam) . 
Curve load factor, 850. 
Custodis chimney, 315. 
Cut-off, point of, 990. 
Cycle, carnot, 971. 

Clausius, 977. 

conventional, 982. 

Rankine, 977. 

rectangular, 982. 

regenerative, 977. 
Cylinder condensation, 366. 
Cylinder cups, 795. 
Cylinder efficiency, 364. 
Cylinder lubrication, 794. 
Cylinder ratio, compound engines, 393. 

Dalton's laws, 502. 
Damper, 8. 

Kitt's hydraulic, 179. 

loss of draft through, 288. 

Tilden steam actuated, 179. 
Daily load curve, 885. 
Dean vacuum pump, 652. 
De Laval centrifugal pump, 668. 

steam turbines, 429, 443. 
Delray station, boiler at, 138. 

boiler test, 160. 
Demand factors, 850-3. 
Density of air and flue gas, 280. 
Depreciation, rate of, 861-7. 
Diagram factor, 984. 
Diamond soot blower, 182. 
Diaphragm valve, 753. 
Differential traps, 696. 



" Direct " steam separator, 684. 
Directly fired superheater, 233. 
Disk-valve pump, 627. 
Disk water meter, 810. 
Distillation of feed water, 606. 
Divergent nozzle, 433. 
Diversity factor, 850. 
Domes on steam boilers, 128. 
Double-arch bridge-wall, 196. 
Double-flow steam turbine, 481. 
Down-draft furnace, 198. 
Down spout, 13. 
Draft, available, 284. 

balanced, 336. 

chimney, 280. 

chimney vs. mechanical, 347. 

combined chimney and forced, 335. 

fans for, 337. 

for various fuels, 284. 

forced, 330. 

friction loss, 285-7. 

induced, 332. 

influence on boiler rating, 283. 

influence on rate of combustion, 330. 

loss in boilers, 285-7. 

mechanical, 327-348. 

steam jets, 328. 

theoretical intensity of, 281. 
Drainage of jackets and receivers, 698. 
Drains, office buildings, 804. 
Drips, high pressure, 690. 

low pressure, 688. 

removal of oil from, 785. 

under pressure, 699. 

under vacuum, 699. 
Dulong's formula, 44. 
Dummy pistons, 478. 
Dunham steam trap, 794. 
Duplex coal value, 276. 

furnace, 215. 

steam pump, 624. 
Dutch ovens, 192. 
Duty, pump, 635. 
Dynamic pressure, 338. 
Dynamometers, 833. 

Economical loading of boilers, 168. 
Economizers (fuel), 607-616. 

coefficient of heat transfer, 612. 

Green, 608. 

heat transmission in, 611. 



1040 



INDEX 



Economizers, proportions in modern 
stations, 613. 
pressure drop through, 610. 
tests of, 615. 
value of, 614. 
Edwards air pump, 654. 
Efficiencies, efficiency (see names of ap- 
paratus in question). 
Ejector condenser, 516. 
Ejector ash, 270. 

Shone, 704. 
Electrical power, cost of, 845-890 (see 

Power costs) . 
Elementary power plants, 1-20. 
piston engine, condensing, 7. 
non-condensing, 7. 
heat and power, 5. 
turbo-alternator, 10-20. 
Elementary theory (see name of appa- 
ratus). 
Elevating tower, 266. 
Ellison calorimeter, 841. 
Emergency governors (see name of appa- 
ratus in question) . 
valves, 765. 
Emulsion tests of oils, 787. 
Engines (steam), 352-423. 

A.S.M.E. Code for, 1007-1020. 
binary vapor, 398. 
Buckeye-mobile, 410. 
classification of, 400. 
clearance volume, 370. 
compound, 405. 

cylinder dimensions of, 393. 
jacketing, 390. 
Lentz, 387. 
Manhattan type, 378. 
non-condensing, 407. 
performance curves, 407. 
receiver for, 391. 
Sulzer, 373. 

superheated steam performance, 421. 
table of best performance, 418. 
terminal pressure in, 372. 
compounding, reason for, 392. 
compression, effect of, 371. 
condensers for, 499-565 (see Con- 
densers) . 
condensing, economy of, 383. 
«ost of, 416. 
cut-off, most economical, 401. 



Engines, cylinder condensation, loss by, 
366. 
cylinder efficiency, 364. 
decreasing back pressure in, 383. 
economic performance of, 417-21. 
economy at various loads, 388-420. 
effect of moisture on economy, 375. 
efficiencies of, 360-5. 
exhaust, loss in, 376. 
Fitchburg-Prosser, 405. 
friction in, 375. 
heat consumption of, 358. 
heat losses in, 366. 
Herrick rotary, 412. 
high speed, 400-4. 
ideal cycles for, 353, 971. 
initial condensation, 366. 
increasing back pressure, 374. 
increasing economy, methods of, 380. 
increasing initial pressure, 380. 
increasing rotative speed, 382. 
indicated horsepower, 357. 
indicated steam consumption, 357. 
intermediate reheating, 391. 
jacketing, 390. 
leakage loss, 368. 
Lentz, 387. 
locomobile, 408. 
logarithmic diagram for, 368. 
losses in, 366. 

mean effective pressure, 984. 
mechanical efficiency, 362. 
medium speed, 404. 
non-condensing, 400-4. 
Nordberg uniflow, 394. 
pressure, increasing initial in, 380. 
pumping, 408. 
quadruple expansion, 408. 
radiation losses in, 376. 
Rankine cycle for, 353, 977. 
Rankine cycle ratio, 363. 
receivers for, 391. 
rotary, 410. 
selection of type, 415. 
simple, 400. 
Skinner uniflow, 395. 
steam consumption, 357-422. 
superheated steam, 386. 
thermal efficiency of, 360. 
thermodynamics of, 971-991. 
throttling vs. automatic cut off, 413. 



INDEX 



1041 



Engines, triple-expansion, 408-9. 

types of, 400. 

uniflow engines, 393-8. 
C & G Cooper, 394. 
Nordberg, 394. 
performance of, 395-7. 
Skinner, 395. 

water rate, 357-422. 

Willans line, 357. 

wire drawing, 374. 
Entropy, 951. 
Equation of pipes, 747. 
Essex station, 914. 
Evaporation, 143. 

cooling pond, 558. 

factors of, 143. 

latent heat of, 945. 

rate of, boilers, 149. 

tests of oils, 787. 

total heat of, 945. 

unit of, 143. 
Evas^ stack, 334. 
Exciters, 12. 

Exhauster steam jet, 272. 
Exhaust head, 5. 
Exhaust steam, heat loses in, 376. 

heating plant, 433. 
Expanding nozzle, 433. 
Expansion, loss due incomplete, 371. 

pipe materials, 705. 

ratio of, 934. 

steam, 960-7. 
Expectancy, 876. 
Extra-strong pipe, 807. 

Factor of evaporation, 143. 
Fan draft, 330. 
cost of, 349. 
Fans (centrifugal), 330-348. 
balance draft, 335. 
Buffalo conoidal, 337-342. 

planoidal, 342. 
capacities of, 346. 
characteristics, 340-3. 
efficiencies of, 349. 

manometric, 344. 

mechanical, 344. 

volumetric, 344. 
forced draft, 332. 
horsepower of, 341-3. 
induced draft, 333. 



Fans, planoidal, 342. 
performance of, 338. 
pressures in, 338-9. 
dynamic, 338. 
static, 338. 
velocity, 339. 
selection of, 345. 
Sirocco, 337-347. 
steel plate, 337, 340. 
turbo undergrate, 334. 
types of, 336. 
Feed water, 566. 
analysis of, 567. 
boiler compounds for, 572. 
chemical purification of, 572-583. 
distillation of, 606. 
economy of pre-heating, 584. 
foaming and priming caused by, 569. 
general treatment of, 571. 
hardness measure of, 566. 
heaters, 585-600. 
atmospheric, 586. 
Baragwanath, 593. 
choice of, 616. 
classification of, 585. 
closed, 590-605. 
Baragwanath, 593. 
coefficient of heat transfer, 596. 
coil, 592. 
economizers, 607. 
film, 594. 
Goubert, 591. 
Harrisburg, 592. 
heat transmission, 594-7. 
multi-flow, 592. 
Otis, 593. 

parallel current, 614. 
single-flow, 591. 
steam tube, 593. 
temperature gradient in, 595. 
types of, 590. 
Wainwright, 592. 
water tube, 590. 
Cochrane, 587. 

coefficient of heat transfer in, 596. 
coil, 592. 

economizers, 607 (see Economizers). 
exhaust steam, 376-379. 
film, 594. 
flue gas, 607. 
Goubert, 591. 



1042 



INDEX 



Feed water, heaters, Harrisburg, 594. 
heat transmission in, 594-7. 
Hoppes, 340. 
induced, 586. 
Hve steam, 604. 
multi-flow, 592. 
open, 586-590. 
atmospheric, 586, 
Cochrane, 587. 
Hoppes, 340. 
induced, 586. 
live steam, 604. 
pan surface for, 589. 
primary, 586. 
secondary, 586. 
size of shell, 589. 
temperature rise in, 588. 
through, 585. 
vacuum, 585. 
vs. closed, 600. 
primary, 585. 
secondary, 586. 
single-flow, 591. 
steam tube, 593. 
impurities in, 255. 
internal corrosion caused by, 570. 
mechanical purification of, 574. 
permanent hardness of, 566. 
piping, 754. 

purifiers and softeners, 577-581. 
Anderson system, 581. 
chemicals for, 575. 
cold process, 577. 
cost of, 582. 
hot process, 577. 
Kennicott system, 578. 
Permutit system, 582. 
Scaife system, 579. 
We-fu-go, 580. 
regulators, 774. 
scale produced by, 568. 
soap solution for testing hardness, 

566. 
softening, 577-581. 
temporary hardness, 566. 
thermal purification of, 574. 
weighing of, 806. 
Fery radiation pyrometer, 830. 
Filters, oil, 801. 
Film heaters, 594. 
Fire thickness, 170. 



Fire, temperature of, 58. 
Fire box boiler, 1 18. 
Fire test, oils, 785. 
Fires, banking, 192. 

building, 191. 

cleaning, 191. 
Fisher pump governor, 640. 
Fitchburg-Prosser engine, 640. 
Fittings, pipe, 711-15. 
Fixed carbon, 29. 
Fixed charges, 858. 
Flameless combustion, 112. 
Flanges, pipe, 711-15. 
Flash point of oils, 849. 
Fleming Harrisburg engine, 361. 
Flight conveyor, 253. 
Flinn differential trap, 695. 
Float trap, 691. 
Flow, steam, nozzles, 436. 

pipes, 740. 

water-pipes, 740. 
Flue gas analysis, 49, 63, 835. 

apparatus, 835-840. 

Hempel, 836. 

Little, 837. 

Orsat, 835. 

Simmance-Abady, 838. 

Uehling, 839. 

Williams, 836. 

composition of, 49. 

heat loss in, 61. 

heaters, 706. 

temperatures of, 164. 
Fly-wheel pumps, 631. 
Foaming in boilers, 569. 
Foot valves, 755. 
Forced capacities of boilers, 166. 
Forced draft, 330. 
Ford gas-steam plant, 19. 
Foster back pressure valve, 772. 

pressure regulator, 774. 

superheater, 231. 
Foundation bolts, steel stacks, 306. 
Foundation, chimney, 322. 
Fountain, spray, 560. 
Four-valve engine tests, 404. 
Free burning coal, 35.. 

hydrogen, 27. 

oxygen, 51. 
Friction in engines, 375. 

pipe fittings, 745, 759. 



INDEX 



1043 



Friction in pipes, 746, 757. 
Friction tests of oils, 778. 
Fuel, calorimeters, 842. 

cost of, 972. 
Fuel oU, 89-109. 

advantages of, as boiler fuel, 89. 
air requirements for, 92. 
analysis of, 90. 
atomization of, 101. 
Baume scale for, 89. 
boiler feeding systems, 101-5. 
boiler tests with, 93-100. 
burners, 94-7. 
Billow, 97. 
Branch, 96. 
Hammel, 96. 
Kirkwood, 97. 
Korting, 94. 
Peabody, 96. 
tests of, 100. 
Warren, 97. 
chemical properties, 89. 
efficiencies of boilers burning, 91 . 
furnaces, 95-9. 
front feed, 99. 
Hammel, 98. 
Peabody, 99. 
Government specifications for purchas- 
ing, 108. 
heating value and gravity of, 91. 
physical properties of, 89. 
purchase of, 108. 
transportation and storage, 105. 
vs. coal for boiler fuel, 94. 
Fuel ratio, 30. 
Fuel, weighing, 804. 
Fuel, calorific value of, 47. 
classification of, 22. 
gaseous, 109. 
sohd, 22. 
Fuels and combustion, 22-111. 
Furnace efficiency, 155. 

temperature, 58. 
Furnaces, 190-221. 

B. & W. boilers, 129, 122-40. 
Badenhausen boiler, 610. 
bagasse, 220. 
Burke's smokeless, 198. 
Chicago settings, 194-9. 
Delray boiler, 138. 
double-arch bridge-wall, 196. 



Furnaces, down draft, 198. 

Dutch oven, 192. 

efficiency of, 157. 

gaseous fuels, 109. 

green bagasse, 220. 

Hammel oil fired, 98. 

hand-fired, 122-40, 190. 

Hawley down draft, 198. 

Meyer tan-bark, 221. 

Murphy smokeless, 211. 

oil fuel, 95. 

Peabody oil fired, 99. 

powdered coal, 85-7. 

Riley duplex, 215. 

smokeless, 190-210. 

steam jets for, 200. 

stoker-fired, 201-218. 

tan bark, 221. 

twin-fire, 194-5. 

waste-heat, 109. 

wing-wall, 197. 

wood refuse, 193. 
Fusible plugs, 181. 
Fusibility of ash, 77. 

Gas burner, Gwynne, 110. 
Gate valves, 763. 
Gauge, pressure, 825. 

water-level, 141, 180. 
Gebhardt steam meters, 815. 
G-E steam meters, 817. 
Geipel traps, 694. 
Globe valves, 762. 
Going value, 868. 

Goodenough, properties of steam, 942-9. 
Goubert feed- water heater, 591. 
Government specification, coal, 909. 

fuel oil, 108. 
Governors {see name of apparatus in 

question). 
Grashof's formula, 436. 
Grate bars, 175. 

efficiency, 154. 

surface, 151. 
Grates, fuel loss through, 65. 

rocking, 175. 

stationary, 175. 

traveling, 203. 
Greases, 780. 
Grease extractor, 686. 
Green bagasse furnace, 220. 



1044 



INDEX 



Green chain grate, 208. 

economizer, 607. 
Gumming tests, oils, 783. 
Gutermuth valves, 655. 
Guyed stacks, 299. 
Gwynne gas burner, 1 10. 

Hammel fuel-oil burner, 96. 

furnace, 98. 
Hammer type tube cleaner, 183. 
Hammler-Eddy smoke recorder, 219. 
Hancock injector, 647. 
Hand-fired furnaces, 122-140, 190. 
Hand shoveling, 251. 
Hangers for piping, 728. 
Hardness test, 566. 
Harrisburgh feed-water heater, 594. 
Hartford Boiler Ins. Co. annual report, 
570. 
boiler specifications, 891. 
Hawley down-draft furnace, 199. 
Header, main steam, 734. 
Heat balance, 69, 1011. 
Heat consumption of prime movers, 358, 

492. 
Heat losses, bare pipe, 720. 
combustion of coal, 60-75. 
dry flue gases, 61. 
fuel in ash, 65. 
hydrogen in fuel, 58. 
incomplete combustion, 62. 
inherent, 73. 
moisture in air, 66. 
moisture in fuel, 67. 
radiation, 68. 
preventable, 74. 
visible smoke, 68. 
covered pipe, 721. 
Heat transmission, boilers, 144-7. 
condensers, 529. 
economizers, 611. 
engines, 366. 
feed-water heaters, 594. 
piping, 720. 
superheaters, 238. 
power plants, 4-20. 
Heaters, feed water, 585-600 (see Feed 

water, heaters). 
Heating surface, boiler, 148. 
Heating systems, 747-751. 
Paul, 750. 



Heating systems, Webster, 748. 
Heating value of fuels, 44. 
Height of chimneys, 298. 
Heine boiler, 130. 

superheater, 232. 
Heintz expansion traps, 695. 
Hempel pipette, 836. 
Herrick rotary engine, 412. 
Herringbone grate, 175. 
High pressures (boiler), 139. 
High-pressure drips, 690. 
High-speed engines, 401. 
High-water alarm, 180. 
High- vacuum condensers, 512. 

pumps, 655. 
Hoist and trolley, 268. 
Holly loop, 702. 
Hoppes feed-water heater, 340. 

steam separator, 681. 
Horsepower, boiler, 150. 
Horizontal tubular boilers, 122. 
Hot-well depression, 534. 

pumps, 675. 

temperatures, 521, 538. 
Humidity, relative, 998. 
Hunt coal conveyor, 262. 
Hurling water, 18. 

pumps, 660. 
Hydraulic air pump, 660. 

governors, 459. 

packing, 630. 
Hydrogen, available, 27. 

combustion data, 43. 

determination of, in coal, 27. 

free, 27. 

heat loss, coal, 68. 

net heating value, 44. 

total, 44. 
Hydrometer, Baum6, 89. 
Hydrostatic lubricator, 796. 
Hygrometry, 998. 

Ideal engine cycle, 353. 

Ideal feed-water temperature, 358. 

Illinois chain grate, 204. 

Impellers for centrifugal pumps, 664. 

fans, 337. 
Impurities in feed water, 255. 
Impulse turbines, 425-446. 
Inadequacy, 861. 
Incomplete combustion, 62. 



INDEX 



1045 



Incomplete expansion, 371, 
Increasing back pressure, 374. 
Independently-fired superheater, 233. 
Indicated horsepower, 357. 
Indirectly-fired superheater, 235. 
Induced draft, 334. 

heaters, 586. 
Inherent furnace losses, 73. 
Initial condensation, 366. 
Injection orifice, 511. 

water, 520. 
Injectors (steam), 645-7. 

automatic, 647. 

performance of, 648. 

positive, 646. 

range in working pressures, 649. 

vs. steam pumps, 650. 
Insurance (power cost), 869. 
Interest charges, 860. 
Intermediate reheating, 391. 
Intermittent oil feed, 789. 
Internal corrosion, 570. 
Isolated stations, cost of power, 860-879. 

W. H. McElwain Co., 929. 
Isothermal change of state, 963. 

Jackets, steam engine, 390. 
Jet condensers, 512-523 {see Condensers, 
jet). 

pumps, 661, 675. 
Jets, steam, 200, 272, 328. 

kinetic energy of steam, 435. 

velocity of steam, 433. 
Jones underfeed stoker, 212. 

Kennicot feed- water purifier, 578. 

water weigher, 807. 
Kents' chimney formula, 296. 
Kerosene in boilers, 573. 
Kerr turbine, 448. 
Kewanee boiler, 118. 
Keystone steam separator, 682. 
Kieley reducing valve, 774. 
Kindhng temperature of fuels, 41. 
Kirkwood fuel-oil burner, 94. 
Kitts hydraulic damper, 179. 
Koerting jet condenser, 517. 
Korting fuel-oil burner, 94. 
Knowles electric geared pump, 643. 

Labor, cost of, in power plants, 871. 
Labyrinth packing, 669. 



Lea Degen pump, test of, 669. 
Leakage of air in condensers, 504, 655. 

steam in engines, 369. 
Leblanc air pump, 659. 
Lentz superheated steam engine, 387. 
Leyland cylinder cup, 790. 
Life of power-plant appliances, 865. 
Lignite, 37. 

"Little" flue-gas apparatus, 837. 
Live-steam feed-water heaters, 604. 
separators, 679-688 (see Separators, 
steam) . 
Load curves, 886. 

factors, 851. 
Locomobile, 408. 
Loew grease extractor, 686. 
Logarithmic diagram, 368, 985. 

mean temperature difference, 507. 
Loop heater, 735. 
Loop, Holly, 702. 

steam, 701. 
Loss of heat, bare pipes, 720. 
combustion of fuels, 60-75. 
covered pipe, 721. 
steam engines, 366. 
Losses standby, 71. 
Low-level jet condenser, 514. 
Low-pressure drips, 688. 

turbines, 479. 
Low-speed engines, 404. 
Low-water alarm, 180. 
Lubricants, 777-787. 
animal fats, 777. 
chemical tests, 781-3. 
acids, 782. 
alkalies, 782. 
effect of heat, 783. 
gumming, 783. 
insoluble in ether, 782. 
insoluble in gasoline, 782. 
moisture, 782. 
sulphur, 782. 
tarry matter, 783. 
graphite, 779. 
greases, 780. 
mineral oils, 778. 
physical characteristics of, 788. 
physical tests, 783-5. 
cold, 785. 
color, 784. 
emulsion, 787. 



1046 



INDEX 



Lubricants, physical tests, evaporation, 
787. 

flash point, 785. 

fire, 785. 

friction, 778, 787. 

gravity, 784. 

odor, 784. 

viscosity, 786. 
quaUfications of good, 781. 
soHd, 770. 
testing, 781. 
vegetable oils, 777. 
service tests, 787. 
Lubrication, 789-802. 

atmospheric surface, 789-4,. 

centrifugal oiler, 791. 

compressed air feed. 794. 

intermittent feed, 789. 

gravity oil feed, 792. 

low-pressure gravity feed, 793. 

oil bath, 790. 

oil cups, 790. 

pendulum oiler, 791. 

restricted feed, 789. 

ring oiler, 791. 

splash, 792. 

telescopic oiler, 790. 
central systems, 798-800. 

Curtis turbine, 799-800. 

piston engine plant, 798. 
cost of, 877. 
cylinder, 794-7. 

cylinder cups, 794. 

hydrostatic, 795. 

forced feed, 797. 
Ludlow angle valve, 764. 
Lunkenheimer lubricator, 796. 



McDaniel float trap, 691. 
Mean specific heat, air, 991. 

gases, 60. 

superheated steam, 949. 

water, 945. 
Mean temperature difference, arithmetic, 
537. 

logarithmic, 536. 
Mechanical boiler tube cleaners, 183. 

draft, 327-350 (see Fans). 

efficiency of fans, 344. 
engines, 362. 
pumps, 632. 

purification of feed water, 574-580. 

stokers, 201-216 (see Stokers). 
Meters, steam, 813-824. 

Bailey, 824. 

classification of, 814. 

Gebhardt, 814. 

G-E, 817. 

Republic, 820. 

St. Johns, 823. 
Methane, combustion data, 43. 
Meyer's tanbark furnace, 221. 
Mineral oils, 778. 
Missing quantity, 366. 
Mixed-pressure turbines, 479. 
Moisture, air, 994. 

combined, 27. 

fuel, 26. 

steam, 679, 943. 

total, 27. 
Mollier diagram, 956. 
Multi-stage centrifugal pumps, 666. 

steam turbines, 425, 453. 
Murphy furnace, 211. 
Myriawatt, def., 1031. 



Mahler bomb calorimeter, 842. 

Main exciter, 12. 

Mains, steam, 734. 

Maintenance, 869. 

Manometric efficiency, 344. 

Marks and Davis' steam tables, 953. 

Marsh boiler feed pump, 628. 

steam pump test, 633. 
Materials, brick chimneys, 312. 

pipes and fittings, 705. 

superheaters, 236. 
Maximum demand, 852. 
McClave's argand blower, 166. 



Napier's rule, flow of steam, 771. 
Natural draft, chimney, 280. 

cooling tower, 501. 
Natural gas, properties of, 109. 
Nitrogen, in air, 47. 

in coal, 47. 

properties of, 43.- 
Non-condensing, engine tests, 417, 422. 

plants, elementary, 2. 

exhaust heating, Paul system, 750. 
Webster system, 748. 

feed-water heating, 754. 
Non-return valves, 765. 



INDEX 



1047 



Nordberg uniflow engine, 394. 
Nozzles, divergent, 433. 

expanding, 433. 

expansion ratio, 436. 

flow of steam through, 435. 
water through, 758. 

friction in, 437. 

mouth error of, 436. 

steam turbine {see name of turbine). 
Nugent's telescopic oiler, 790. 

Obsolescence, 861. 
Oil, burners, 96-99. 

fuel, 89-109 (see Fuel oil). 

piping, 798. 

relay governor, 451. 

separators, 685. 
Oiling systems, 789-802 (see Lubrication). 
Oils, 777-787 (see Lubricants). 
Open heater vs. closed, 600. 
Open heaters, 596-590. 
Open-top conveyor, 256. 
Operating costs, 869. 
Optical pyrometer, 829. 
Orifice measurements, 812. 

size of injection, 511. 
Orifices, flow of steam through, 435. 

flow of water through, 757. 
Orrok, tests by, 534-7. 
Orsat apparatus, 835. 
Otis heater, 593. 
Output and load factor, 849. 
Ovens, Dutch, 192. 
Overfeed stokers, 207. 
Overhead charges, 858. 
Overload capacity, boilers, 163. 
Oxygen, in air, 47. 

in coal, 23. 

in flue gases, 51. 

properties of, 43. 
Oxygen-hydrogen ratio, 31. 

Packing, pump, 630. 

Pacific Light & Elec. Co., power cost, 881 . 

Pan conveyor, 256. 

surface, open heater, 589. 
Parallel current condenser, 508. 
Parallel flow, economizers, 612. 
Parker boiler, 132. 
Parr coal calorimeter, 842. 
Parsons vacuum augmenter, 661. 



Paul exhauster, 751. 

steam heating system, 751. 
Peabody oil burner, 94. 

oil furnace, 99. 
Peat, 37. 

Peck conveyor, 258. 
Penberthy injector, 647. 
Pendulum oiler, 791. 
Pennel evaporative condenser, 543. 
Permanent statistics, power plant, 847. 
Permutit feed-water purification, 583. 
Pipe, pipes, 705-720. 
anchors, 728. 
and fittings, 705-720. 
bends, 725. 

expansion, 725. 

minimum dimension, 726. 
boiler tubes, 711. 
brass, 707. 

bursting strength of, 708. 
capacity per foot length, 710. 
cast-iron, 706-713. 
cast-steel, 707. 
copper, 707. 
coverings, 721. 

heat loss through, 721. 
drains, 688. 
drawings, 705. 
equations, 747. 
expansion, 724. 
flanges, 711-715. 

American Standard, 715. 

cost of, 719. 

types of, 712. 
fittings, 711-715. 
flow of steam in, 740. 
flow of water in, 757. 

American Standard, 715. 

flanged, 711. 

resistance, steam flow, 746. 

resistance, water flow, 757. 

screwed, 709. 
friction in, 745, 758. 
heat conduction through, 720. 
materials for, 705. 
pressure drop in, 746, 757. 
riveted, 708. 
screw threads for, 709. 
size of, 707-710. 
steel, 706-10. 
supports, 728. 



1048 



INDEX 



Pipe, Van Stone joint for, 713. 
welded joints for, 718. 
wrought iron, 706-9. 
bursting pressure, 708. 
corrosion of, 707. 
double extra strong, 708. 
extra strong, 708. 
large O. D., 708. 
standard, 708. 
Piping, 705-751. 
blow-off, 769. 
condenser, 548-554. 
duplicate system, 731. 
exhaust steam, 747. 
feed water, 754. 
heating systems, 748-751. 
Paul, 750. 
Webster, 748. 
high-pressure steam, 730-742. 
oil, 798-800. 
loop in ring header, 735. 
single header, 733. 
spider system, 732. 
steam, examples of, 731-742. 
Pistons, water, 630. 
Pitot tubes, 338. 
Pivoted-bucket conveyors, 260. 
Poly tropic change of state, 967. 
Ponds, cooling, 558. 
Pop safety valves, 770. 
Poppet-valve engine, 390. 
Positive injectors, 646. 
Powdered coal, 80-88. 
advantages of, 80. 
burners, 81-84. 
Blake, 84. 
forced draft, 82. 
Mann, 83. 
natural draft, 82. 
Rowe, 82. 

United Combustion Co., 83. 
cost of preparing, 87. 
fineness for boiler fuel, 87. 
furnaces, 84-7. 

Am. Locomotive Co., 86. 
depreciation of, 88. 
efficiency of, 88. 
Mann, 85. 

M. K. & T. shops, 87. 
storing, 87. 
Power costs, 845-890. 



Power costs, amortization, 868. 
attendance, 871. 
apartment building, 879. 
bibliography, 888. 
central stations, 870-881. 

Commonwealth Edison, 870. 

Delray Power House, 857. 

Fort Wayne Municipal, 877. 

Illinois, 873. 

Iowa statistics, 882. 

Massachusetts, 872, 875. 

Pacific Light & Electric, 881. 
curve load factor, 850. 
demand factors, 850-3. 
depreciation, 861-7. 
diversity factor, 850. 
expectancy, 867. 
flat rates, 849. 
fixed charges, 858. 
fuel costs, 872. 
general remarks, 857. 
going value, 868. 
inadequacy, 861. 
insurance, 869. 
interest, 860. 
isolated stations, 878-9. 

initial cost, 860. 
labor, 871. 

life of equipment, 865. 
load factors, 851. 
locomobile plants, 876. 
maintenance, 869. 
manufacturing plants, 884. 
maximum demand, 852. 
obsolescence, 861. 
oil, waste, etc., 877. 
operating costs, 869. 

records, 847. 
output and load factor, 849. 
permanent statistics, 847. 
piston engine plants, 880. 
records, 845. 
repairs, 877. 
sinking fund, 866. 
station load factor, 850. 
steam heating, 878. 
straight line depreciation, 863. 
taxes, 869. 

turbo-electric plants, 873-880. 
unit rate, 849. 
Power driven pumps, 644. 



INDEX 



1049 



Power measurements, 832. 
Power plant design, elements of, 881. 
Power plants, A.S.M.E., testing codes, 
1023. 
elementary, 1-20. 
transmission losses, 20. 
typical central stations, 914-928. 
typical isolated stations, 929-991. 
Powers thermostat, 753. 
Pressure drops in, boilers, 286-9. 
condensers, 526. 
economizers, 612. 
piping, steam, 744, 746. 
water, 758. 
Pressure regulator, 774. 
Pressures, high boiler, 380. 
Preventable combustion losses, 74. 
Primary heaters, 586. 
Products of combustion, 49. 
Properties of, air, 991-1006 (see Air, 
properties of). 
fuel oil, 90. 
gases. 111. 
lubricating oil, 788. 
steam, 942-971 (see Steam, properties 
of). 
Proximate analysis of, 23. 
Psychrometric chart, 996. 
Pulsometer, 674. 
Pump governor, 639. 
Pumping engine tests, 409. 
Pumps, 622. 

air, 651-662 (see Pumps, vacuum). 
air chambers for, 629. 
air lift, 676. 
boUer-feed, 624-50. 
centrifugal, 674. 
direct-acting, 624. 
duplex, 624. 
Marsh, 628. 
simplex, 628. 
size of, 638. 

steam-end, duplex, 626. 
tests of, 632-3. 
valve-gear, duplex, 626. 
classification of, 622. 
centrifugal, 664-9. 
boiler-feed, 674. 
characteristics of, 668-74. 
circulating, 672. 
condensate, 675. 



Pumps, centrifugal, hot-well, 675. 

impellers for, 664. 

Lea-Degen, test of, 669. 

multi-stage, 666. 

performance of, 668-74. 

power requirements, 673. 

Rees-roturbo, 653. 

single stage, 665. 

turbine, 666. 

vacuum service, 651-62. 

volute, 669. 

Worthington, 3-stage, 666. 
circulating, 672, 
condensate, 675. 
condenser, 651-662. 
Dean wet-vacuum, 652. 
dry vacuum, 658. 
duplex piston, 624. 
duty of, 635. 
Edwards air, 654. 
efficiencies of, 639. 
electric driven piston, 643. 
fly wheel, 631. 
geared piston, 644. 
governors for, 639. 
high vacuum, 655. 
hot- well, 675. 
hurling water, 660. 
hydraulic air, 660. 
injector, 645-7 (see Injectors). 
jet, 622, 674. 

Knowles electric, piston, 643. 
Leblanc, air, 659. 
outside packed plunger, 630. 
packing for piston, 630. 
performance of, 632. 
plungers for, 630. 
power requirements of, 633-74. 
pulsometer, 674. 
reciprocating piston, 624-650. 

air, 651 (see Pumps, vacuum). 

air chambers for, 629. 

boiler-feed, 624-30. 

compound duplex, 627. 

duplex, 624. 

duty, 635. 

efficiencies, 632. 

electric driven, 643. 

fly-wheel, 631. 

geared, 644. 

governors, 639. 



1050 



INDEX 



Pumps, reciprocating piston, Knowles 
electric, 633. 

Marsh simple, 628. 

outside packed, 631. 

performance of, 632. 

piston packing, 630. 

plungers, 630. 

power driven, 644. 

simplex, 628. 

size of boiler-feed, 638. 

slip, 638. 

steam end, 626. 

tests of, 632. 

water end, 630. 
Rees-roturbo, 653. 
screw, 672. 
rotary, 670. 
rotrex, 655. 
simplex steam, 628. 
slip in, 638. 
steam, 624-40. 
tail, 657. 
tests of, 632. 
Thyssen, 659. 
triplex, 644. 
turbo-air, 660. 
vacuum, 651-662. 

air handled by, 656. 

centrifugal-hydraulic, 659. 

Dean, 652. 

dry, 658. 

Edwards, 654. 

hurling water, 659. 

hydraulic, 659. 

Leblanc, 659. 

performance of, 663. 

radojet, 661. 

Rees-roturbo, 653. 

rotrex, 655. 

size of, 655. 

Thyssen, 659. 

turbo-air, 660. 

wet, 651. 

Wheeler, 658-660. 

Worthington, 659. 
wet-vacuum, 651. 
Wheeler rotrex, 655. 

turbo-air, 660. 
Worthington hydraulic vacuum, 660. 

two-stage vacuum, 658. 
Purchasing coal, 75, 909. 



Purchasing fuel oil, 108. 
Purification, feed-water, 574-583. 
Purifiers, live steam, 604, 
Purifying plants, feed-water, 574-583. 
Pyrometers, air, 828. 

optical, 829. 

radiation, 830. 

resistance, 829. 

thermo-electric, 828. 

Quadruple-expansion engines, 408. 
Quality of steam, 942. 

Radial brick chimneys, 317. 
Radial flow condensers, 526. 
Radiation, heat transmitted by, 144. 
Radiation losses, boilers, 68. 

engines, 376. 

pyrometers, 800. 
Radius of statical moment, 313. 
Radojet pump, 661. 
Rankine cycles, 363, 977-982. 
Rate of combustion, 151. 

depreciation, 861. 

driving boilers, 163. 
Rating of boilers, 150. 
Ratio of expansion, 393. 
Rateau regenerator-accumulator, 483. 

installations, 482. 
Reaction turbines, 467, 481. 
Receiver reheaters, 391. 
Reciprocating engines, 624, 650 (see 

Engines, steam). 
Records, power plant, 845. 
Recording apparatus {see name of ap- 
paratus in question) . 
Redondo plant, overall, efficiency, 10. 
Reduction gears, 446. 
Rees-Roturbo pump, 653. 
Regenerator accumulator, 483. 
Regulators, damper, 179. 
Reinforced concrete chimneys, 317. 

feed water, 774. 
Relief valves, 777. 
Repairs, cost of, 877. 
Republic flow meter, 820. 
Return traps, 697. 
Return-tubular boilers, 122. 

specifications for, 891. 
Returns tank, 703. 

Rhode Island power-house, ash-handling, 
263. 



INDEX 



1051 



Riley stokers, 166, 214. 
Ring oilers, 791. 
Ring steam jet, 328. 
Ringelmann smoke chart, 217. 
Riveted joints, chimney, 305. 
Robb-Mumford boilers, 121. 
Robins belt conveyor, 264. 
Rochester forced-feed lubricator, 797. 
Roney stoker, 207. 
Rotary engines, 410. 

pumps, 670. 
Rotrex air pumps, 655. 
Rowe coal dust feeder, 821. 

feed-water regulator, 82. 

Safety plugs, 181. 
Safety valves, 770. 

capacity of, 771. 

dead weight, 770. 

lever, 770. 

pop, 770. 
Saturated air, properties of, 994. 

surface condensers, 543. 
Saturated steam, properties of, 942-960 

(see Steam, properties of). 
Sawdust as fuel, 38. 
Scaife feed-water purifier, 579. 
Scale, analysis of boiler, 567. 

influence on heat transmission, 569. 

methods for removing, 574. 
"S-C" feed-water regulator, 642. 
Schmidt independent superheater, 234. 
Schutte ejector condenser, 517. 
Schwoerer superheater, 231, 245. 
Scioto Valley Traction Co., coal-handling 

system, 269. 
Scotch marine boiler, 119. 
Scraper conveyors, 252. 
Screw conveyors, 252. 

pump, 672. 
Screwed fittings, 709. 
Seaton coal valve, 277. 
Secondary heaters, 586. 
Seger cones, 831. 
Semi-anthracite coals, 33. 

bituminous coals, 34. 
Separating calorimeters, 841 
Separators, 679-688. 

live steam, 679-686. 
Austin, 683. 
Bundy, 682. 



Separators, live stoam, classification of, 
680. 
"Direct," 684. 
efficiency of, 680. 
exhaust heads, 687. 
Hoppes, 681. 
Keystone, 682, 
location of, 684. 
Stratton, 681. 
tests of, 680. 
types of, 680. 
oil, 685-7. 
Baum, 685. 
Loew, 686. 
Service tests, lubricants, 787. 
Settings, boiler, 122, 190-221. 
Shaking grates, 177. 
Shone ejector, 704. 
Side feed stokers, 211. 
Simmance-Abady flue-gas recorder, 838. 
Simplex coal valve, 276. 

steam pumps, 638. 
Sinking fund, 860. 
Single-stage pumps, 665. 

turbines, 429. 
Siphon condensers, 515. 

traps, 696. 
Sirocco blowers, 337. 
Skimmer, Buckeye, 178. 
Skinner unifiow engine. 395. 
Slip expansion joint, 728. 
Slip in pumps, 638. 
Smoke, cause of visible, 189. 
chart, 217. 

chemical elements in visible, 188. 
determination of, 217. 
distribution in Chicago, 190. 
Hammler-Eddy recorder, 219. 
loss due to visible, 68, 187. 
prevention, 187-216. 
recorder, 219. 
Ringelmann chart, 217. 
solids in visible, 188. 
units, 219. 
Smokeless furnaces, 184-9 (see Furnaces). 
Soap, standard solution, 566. 
Softeners, feed-water, 577-581. 
Solid lubricants, 770. 
Solids in visible smoke, 188. 
Soot blowers, 181. 
Soot, loss due to, 69. 



1052 



INDEX 



Space, air, in boiler settings, 127. 

in grate bars, 176. 
Specific gravity, Baum6, 784. 

fuel oils, 784. 

lubricating oils, 89. 
Specific heat, air, 991. 

gases, 60. 

saturated steam, 997. 

superheated steam, 947. 

water, 945. 
Specific volume, saturated steam, 943. 

superheated steam, 943. 
Specifications, typical boiler, 891. 

Government coal purchase, 910. 

piping for central station, 898. 
Speed, economy of increasing, 382. 

measurements of, 832. 
Spider system of piping, 732. 
"Spiro" turbine, 499. 
Splash lubrication, 792. 
Spontaneous combustion, 250. 
Spray fountain, 559. 
Sprinkling stokers, 216. 
Stability, brick chimneys, 313. 

steel chimneys, 306. 
Stacks, 279-326 (see Chimneys). 
Standby losses in boilers, 71. 
Station load factor, 878. 
Steam, boilers, 115-185 (see Boilers, 
steam). 

calorimeters, 840-843. 

condensers, 499-565 {see Condensers). 

consumption of engines, 417-424. 
pumps, 632. 
turbines, 488-490. 

domes on boilers, 128. 

electric power plants, power cost, 845- 
890. 

engines, 352-423 (see Engines, steam). 

flow in pipes, 740. 

through divergent nozzles, 433. 
nozzles, 436. 
orifices, 435. 

gauges, 825. 

jet blower, 328. 

jets, 200, 328. 

loop, 702. 

mains, 734. 

piping, 705-750. 

properties of saturated and super- 
heated, 942-960. 



Steam, properties of saturated and 
superheated, entropy, 951. 
heat of liquid, 944. 
latent heat, 946. 
Mollier diagram, 659. 
notations, 942. 
quality, 943. 
specific heats, 947. 
standard units, 942. 
tables, saturated, 953. 

superheated, 957. 
temperature-pressure, 943. 
pumps, 624-640 (see Pumps). 
separators, 679-688 (see Separators). 
traps, 690-700 (see Traps, steam). 
tube heaters, 593. 

turbines, 424-500 (see Turbines, 
steam) . 
Steel chimneys, 299-308. 

concrete chimneys, 317-321. 
Stirling boiler, 137. 
superheater, 230. 
St. John's steam meter, 823. 
Stokers, 201-216. 

Babcock & Wilcox, 204. 
chain grates, 203. 
cost of, 222. 
front feed, 207. 
Green, 207. 
Illinois, 204. 
Jones, 212. 
mechanical, 201. 
Murphy, 211. 
overfeed, 207. 
Riley, 214. 
Roney, 207. 
side-feed, 211. 
sprinkling, 216. 
Swift, 216. 
Taylor, 213. 
underfeed, 210. 
Wilkinson, 209. 
Stop valves, 761. 
Storage, coal, 761. 

fuel oil, 105. 
Straight-flow engines, 393. 

line depreciation, 863. 
Straw, fuel value, 38. 
Stress-strain diagram (Tumeaure and 

Maurer), 320. 
Sub-bituminous coal, 35. 



INDEX 



1053 



Sulphur, combustion data, 43. 
in coal, 32-36. 

in lubricants, determination of, 782. 
in visible smoke, 188. 
Sulzer engine, 373. 
Superheat, advantages of, 223. 
economy of, 225, 386. 
limit of, 226. 
Superheated steam, properties of, 942- 

960. 
Superheaters, 227-248. 
Babcock & Wilcox, 229. 
coefficient of heat transfer in, 239. 
flooding device for, 230. 
Foster, 231. 

heat transmission in, 238. 
Heine, 232. 

independently fired, 235. 
integral, 229. 
maintenance of, 237. 
materials for, 236, 
performance of, 242-8. 
Schmidt, 234. 
Schwoerer, 231, 245. 
Stirling, 230. 

temperature range of gases in, 242. 
types of, 227. 
Superheating moisture in air, loss due 
to, 66. 
surface, materials for, 236. 
extent of, 238. 
Supports, pipe, 728. 
Surface blow, 78. 
Surface combustion, 112. 
Surface condensers, 523-547 {see Con- 
densers). 
Surface, cooling for condensers, 532. 
heating for boilers, 148. 
economizer, 611. 
feed-water heaters, 594. 
superheaters, 238. 
Suspended boiler setting, 125. 
Swift sprinkling stoker, 216. 

Tachometers, 833. 
Tail pipe, 15, 657. 
Tanbark, fuel value, 38. 

furnace, 221. 
Tank, blow off, 177. 

returns, 703. 
Tarry matter in lubricants, 783. 



Taxes, 869. 

Taylor underfeed stoker, 213. 
Telescopic oiler, 790. 
Telpherage, 268. 
Temperature, combustion, 58. 
drop in boilers, 172. 
entropy diagram, 987. 
flue gas, forced ratings, 36, 283. 
gradient in economizers, 612. 

in heaters, 595. 
initial, influence of, 169. 
kindling, fuels, 41. 
measurements, 827. 
pressure, steam, 943. 
regulators, 752. 
Temporary hardness, feed waters, 566. 
Terminal pressures, engines, 372. 
Terry steam turbines, 444. 
Tesla bladeless turbine, 498. 
Tests, air pump, 663. 

A.S.M.E. Codes, boiler, 1007. 
engines, 1012. 
power plants, 1023. 
pumping machinery, 1020. 
turbines, 1017. 
blowers, 342-346. 
boilers, 158-170. 
condensers, 540. 
cooling ponds, 560. 
cooling towers, 563. 
economizers, 615. 
engines, 402-423. 
fuel-oil boilers, 93, 167. 

burners, 100. 
injectors, 648. 
lubricants, 779-788. 
pipe coverings, 721. 
pumps, 633-675. 
separators, 680. 
spray fountain, 560. 
steam jets, 329. 
superheaters, 243-248. 
turbines, 489-490. 
Thermal efficiency, engines, 360. 

purification, feed water, 574. 
Thermodynamics, elementary, 960-999. 
change of state, adiabatic, 964. 
constant heat content, 964. 
equal pressure, 960. 
equal volume, 962. 
isothermal, 963. 



1054 



INDEX 



Thermodynamics, change of state, poly- 
tropic, 967. 
ideal cycles, 971-990. 
Carnot, 971. 
Clausius, 977. 
conventional diagram, 983. 
logarithmic diagram, 985. 
Rankine, 977-982. 
rectangular, 982. 
regenerative, 976. 
temperature-entropy, 987. 
properties of steam, 942-959. 
steam turbine, 426. 
Thermometers, 827. 
Thermostat, powers, 752. 
Thickness of fire, 170. 
brick chimney, 308. 
steel chimney plates, 308. 
Throttling calorimeter, 840. 
moisture evaporated by, 964. 
energy loss due to, 964. 
vs. automatic cut-off, 413. 
Thrust bearing, 463. 
Thyssen vacuum pumps, 659. 
Tilden damper regulator, 179. 
Tomlinson condenser, 520. 
Total moisture in fuel, 27. 
Towers (cooling), Barnard- Wheeler, 561. 
C. H. Wheeler, 562. 
tests of, 563. 
Traps (steam), 690-700. 
bucket, 692. 
Acme, 692. 
classification, 690. 
differential, 695. 
Flinn, 696. 
siphon, 696. 
dumps, 692. 

Bundy, 693. 
expansion, 693-5. 
Columbia, 693. 
Dunham, 694. 
Geipel, 694. 
Heintz, 695. 
float, 691. 

McDaniel, 691. 
location of, 696. 
return, 697. 
Shone ejector, 704 
types of, 691. 



Traveling coal hoppers, 275. 

grates, 203. 
Triple-expansion engines, 408. 
Triplex pumps, 644. 
Try cocks, 3, 180. 
Tube cleaners, 180. 
Turbine pumps, 664. 
Turbine tube cleaners, 180. 
Turbines (steam), 424-500. 
advantages of, 486. 
Allis-Chalmers, 473. 
A.S.M.E., testing code, 1017. 
bleeder, 486. 

bucket friction loss in, 443. 
buckets, velocity diagram, 440-478. 
carbon packings for, 464. 
classification of, 424. 
compound cylinder, 460. 
compound pressure, 425. 
compound velocity, 425. 
cost of, 486. 
Curtis, 453-467. 
assembly of valve gear, 461. 
carbon packing, 464. 
elementary theory, 464. 
emergency governor, 461. 
hydraulic governor, 459. 
main controlling governor, 459. 
main operating governor, 458. 
nozzle and blade arrangement, 455. 
performance of, 490. 
single stage, 454. 
steam belt area, 458. 
steam valve gear, 462. 
30,000-kw. compound cylinder, 460. 
throttling governor, 456. 
thrust bearing, 463. 
12,500-kw. single cylinder, 457. 
velocity diagram, 465. 
water cooled bearing, 463. 
De Laval, 429-443. 

multi-velocity-stage, 448. 
single-stage impulse, 429. 
blade assembly, 430. 
elementary theory, 432. 
nozzle, 431. 

theoretical expanding nozzle, 433. 
velocity diagram, 440. 
double-flow, 481. 
economy of space, 487. 



INDEX 



1055 



Turbines, efficiencies of, 488. 
effect of vacuum on, 496. 
effect of superheat on, 494. 
elementary theory, 426. 
exhaust steam, 479. 
heat consumption of, 492. 
high initial pressure in, 496. 
impulse, 425. 

compound pressure, 448. 

compound velocity, 447. 

design of multi-stage, 464. 

design of single stage, 432. 

nozzle proportions, 436. 
Kerr, 448-453. 

bucket fastening, 450. 

eight-stage assembly, 449. 

elementary theory, 452. 

nozzle and vanes, 450. 

oil relay governor, 451. 
labyrinth packing for, 473. 
low pressure, 479. 
maintenance and attendance, 487. 
mixed pressure, 479. 
multi-stage, 451. 
overload capacity, 488. 
performance of, 490. 
Rankine cycle ratio of, 490. 
reaction, 467-481. 

blading for, 470. 

drum diameter ratio, 468. 

elementary theory of, 477. 

end thrust in, 468. 

grouping of stages, 468. 

Westinghouse, 469. 
reduction gears for, 446. 
regulation, 488. 
" Spiro," 499. 

steam consumption of, 490. 
superheat in, 494. 
Terry, 444-453. 

condensing unit, 453. 

elementary theory, 446. 

non-condensing unit, 444. 

reversing chambers, 445. 
Tesla bladeless, 498. 
tests of, 488-494. 
Westinghouse, 446-486. 

blade arrangement, 467. 

blade fastening, 468. 

bleeder, 486. 



Turbines, Westinghouse, compound cyl- 
inder, 477. 
performance of, 484, 490. 
direct governor control, 472. 
double-flow, 476, 481. 
dummies for, 468. 
elementary theory, 477. 
high-pressure non-condensing, 469. 
impulse, 446. 

reduction gear, 446. 
reversing chamber, 446. 
impulse-reaction, 475. 
• labyrinth packing, 473. 
mixed pressure, 485. 
oil relay governor, 470. 
oil relay valve-gear, 471. 
performance of low-pressure, 484. 
performance of 30,000-kw., 490. 
single-flow reaction, 467. 
Turbo air pumps, 18, 660. 
alternator plants, 10, 914. 
electric plants, power costs, 872-880. 
Turneaure and Maurer, stress diagram, 

320. 
Turner oil filter, 802. 
Twin fire-arch furnace, 194. 

Uehling composimeter, 839. 
Ultimate analysis, 25. 
Underfeed stokers, 210. 
Uniflow engines, 393-8. 

C. & H. Cooper Co., 394. 

Nordberg, 394. 

performance of, 395-7. 

Skinner, 395. 
Unit of evaporation, 143. 
Units, conversion, 1031. 

smoke, 219. 
Universal calorimeter, 841. 
United Combustion Co. coal-dust feeder, 

83. 
Useful life of power plant apparatus, 865. 

Vacua, influence of, on engine economy, 
383. 
influence of, on turbine economy, 496. 
Vacuum ash conveyor, 270. 
augmenter. Parsons, 661. 
chambers, 629. 

corrections for standard conditions, 
501. 



1056 



INDEX 



Vacuum, degree of, as affected by air, 
504. 
as affected by aqueous vapor, 505. 

heaters, 585. 

high condensers, 512, 523. 

pumps, 651-662 {see Pumps, vacuum). 
Valves, angle, 764. 

atmospheric rehef, 773. 

automatic stop, 765. 

back pressure, 772. 

blow-off, 767. 

check, 766. 

diaphragm, 753. 

disk, 775, 776. 

emergency, 765. 

foot, 775. 

gate, 763. 

globe, 762. 

Gutermuth, 655. 

non-return, 765. 

pressure regulating, 774. 

pumps, 627. 

reducing, 774. 

safety, 768. 

stop, 761. 
Van Stone joint, 713. 
V-bucket conveyor, 257. 
Vegetable oils, 777. 

Velocity diagrams {see name of apparatus 
in question). 

through blowers, 341. 
nozzles, 435, 758. 
pipes, 740, 758. 
Venturi meter, 811. 
Vertical tubular boilers, 116. 
Viscosity of oils, 786. 
Visible smoke, loss due to, 68, 187. 
Volatile matter in coal, 24. 
Volumetric efficiency, 344. 
Volute centrifugal pump, 664. 
Vulcan soot cleaner, 182. 

Wainwright fuel-water heater, 592. 
Wall brackets, piping, 729. 
Wanner optical pyrometer, 829. 
Warren fuel-oil burner, 97. 
Washed coal, 79. 
Waste-heat boilers, 109, 336. 

gases, 109. 
Water, alkalinity of, 577. 



Water, analysis, 566. 

acidity, 577. 

columns, 180. 

condensing, 521, 527. 

cooling systems, 558. 

flow of, in pipes, 758. 

friction coefficient in pipes, 759. 

gauges, 141, 180. 

hardness of, 566. 

measurements of, 806. 

meters, 806. 

purifying plants, 577-581. 

rates, prime mover {see name of ap- 
paratus) . 

softening plants, 577-581. 

vapor, thermal properties of, 942- 
959. 
Weathering of coal, 250. 
Weber concrete chimney, 318. 
Webster feed-water heater, 587. 

steam heating system, 750. 

vacuum valve, 749. 
We-Fu-Go purifying system, 580. 
Weighing fuel, 804. 

water, 806. 
Weight of air per pound of various fuels, 
55. 

boiler compound necessary, 575. 

guyed steel chimneys, 299, 300. 
Weir measurements, 809. 
Weiss barometer condenser, 518. 
Welded pipe flanges, 718. 
Westinghouse condenser, 587. 

Leblanc air pumps, 659. 

turbines, 445-486. 
Wet air or vacuum pumps, 651. 
Wheeler condenser, 514-525. 

cooling tower, 561. 
Wheeler, C. H., air pump, 658. 

condenser, 513. 

cooling tower, 562. 
White Star oil filter, 801. 
Wickes boiler, 134. 
Wiederholt chimney, 317. 
Wilcox water weigher, 808. 
Wilkinson stoker, 209. 
Willans line, 357. 
Williams flue-gas apparatus, 836. 
Wind pressure, 302. 
Wing-wall furnace, 197, 



INDEX 



1057 



Winslow boiler, 139. > 
Wire drawing, 374, 964. 
Wood, fuel value of, 39. 

refuse furnace, 193. 
Worthington hydraulic vacuum pump, 
660. 

jet condenser, 509. 

multi-stage centrifugal pump, 666. 



Worthington water meter, 809. 

weight determinator, 807. 
Wrought iron pipe, 706, 707. 



Yonkers power house piping, 736. 
Zinc, use of, in boilers, 573. 



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